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Compressed Air Plant 



THE PRODUCTION, TRANSMISSION AND 

USE OF COMPRESSED AIR, WITH 

SPECIAL REFERENCE TO 

MINE SERVICE 



By ROBERT PEELE 

Mining Engineer and Professor of Mining in the School of Mines, Columbia University 



SECOND EDITION 

REVISED AND ENLARGED 



NEW YORK : 

JOHN WILEY & SONS 

London = CHAPMAN & HALL, Limited 
1910 






^ ^^ 



Copyright, igo8, iqio, 
iY ROBERT PEELE 



PRESS OF THE PUBLISHERS PRINTING COMPANY, NEW YORK, U. S. A.. 



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PREFACE TO THE SECOND EDITION 

This edition has been revised and substantially enlarged. 
Among the principal additions are some 90 pages of text and 
63 illustrations, relating to the construction and operation of 
rock-drills, coal-cutting machines and channeling machines. This 
material is contained in Chapters XX, XXI, XXII, and XXIII. 
The detailed records of work of machine drills, in Chapters XX 
and XXI, I believe, will be found useful. Most of the data has 
not before been in print. 

In Chapter III the theory of the compression of air is presented 
in greater detail, together with its applications to the operation 
and performance of compressors. The deductions of the more 
important formulae are also given, such as those used for calculat- 
ing the horse-power required for single- and multiple-stage com- 
pression. In this connection I desire to acknowledge the kind 
assistance of Professor Charles E. Lucke, of Columbia University, 
and Professor H. J. Thorkelson, of the University of Wisconsin. 
To Dr. Lucke my thanks are due for the use of his valuable, and 
hitherto unpublished, notes relating to the work cycles of air com- 
pression, with and without clearance. I would call attention also 
to the records of compressor tests in the latter part of Chapter X. 
These comprise a few typical tests, selected from a large number 
recently made by Mr. R. L. Webb, Mechanical Engineer, on com- 
pressors of different kinds in a well-known Canadian mining 
district. 

Other new material has also been added, relative to the piston 
clearance of the air cylinders of compressors, and the ratio of inlet 
valve area to cylinder area. Numerous minor additions to the 
text have been made, together with corrections and alterations 
where required. The new matter aggregates some 135 pages of 
text and 87 illustrations. Many of the illustrations have been 



IV PREFACE 

furnished by the respective makers of the machinery, to whom 
credit is duly given. In preparing this revision I have kept in 
mind certain kindly criticisms and suggestions received from 
readers of the first edition. R. P. 

New York, June, 19 lo. 



PREFACE TO THE FIRST EDITION 

The increasing use of compressed air makes the subject of 
interest to practitioners in nearly all branches of engineering. 
Besides its more important power applications, such as the 
operation of rock-drills, air brakes, riveting machines, and rail- 
road switching and signalling systems, the uses of compressed air 
are numerous in many minor branches of mechanical engineer- 
ing, in caisson work and the construction of subaqueous founda- 
tions, and in manufacturing industries, chemical works, etc., 
where it serves a multitude of purposes entirely distinct from that 
of the transmission of power. 

A realization of the breadth of the field has suggested that a 
book may be acceptable, addressed especially to those who are 
engaged in mining, tunnelling, quarrying, and other work in- 
volving the excavation of rock, with its concomitant operations. 
While the literature bearing upon this branch of compressed-air 
service is by no means small, it is for the most part scattered 
through the technical periodicals and transactions of engineering 
societies, and therefore not readily accessible to those who are 
out of convenient reach of engineering libraries. I am aware that 
little that is new can be said on this subject, and in writing the 
book I have availed myself freely of existing sources of informa- 
tion. 

In the first part, I have endeavored to present a view of cur- 
rent practice as to the construction and operation of compressors. 

Portions of the subject are dealt with at some length, for ex- 
ample: the types of compressor suitable for different kinds of 



PREFACE V 

service, heat losses occurring in air compression, and the various 
forms of valves, valve-motions, and governing and unloading 
mechanisms, that constitute prominent features of modern com- 
pressor practice. A brief review is given of a few of the funda- 
mental principles of air compression, but my intention has been 
to present only enough of the theory to make intelligible the for- 
mulae employed for the ordinary calculations of the power and 
capacity of compressed-air plant, together with the questions con- 
cerning temperature changes, as affecting the production and use 
of compressed air. Many details of the design of compressors and 
proportions of their parts have been omitted, since these fall prop- 
erly within the province of the mechanical engineer. 

The second part is devoted to the applications (largely to mine 
service) of compressed-air transmission of power, including ma- 
chine drills, pumps operated by compressed air, and mine haul- 
age by compressed-air locomotives. 

Many of the illustrations are reduced or adapted from work- 
ing drawings kindly supplied by compressor builders. Others 
have been taken from catalogues of compressed-air machinery 
and from technical periodicals and books dealing with the different 
types. The origin of these has been stated in nearly every in- 
stance. My thanks are due to several of the technical journals, 
especially Compressed Air Magazine and Mines and Minerals, 
for many suggestions and in some cases for passages extracted 
either in substance or verbatim, from articles therein contained. 
For any important use or adaptation of published material, 
permission has been asked and obtained, and frequent references 
are given in foot-notes or in the body of the text. I also wish to 
acknowledge my indebtedness to Mr. Frank Richards's book on 
" Compressed Air," from which I have derived substantial as- 
sistance. 



ROBERT PEELE. 



School of Mines, Columbia University, 
New York, May, 1908. 



CONTENTS 



PART FIRST 

PRODUCTION OF COMPRESSED AIR 

PAGE 

Preface, iii 

List of Illustrations, xi 

CHAPTER I 

Introduction. Development of Air Compressors. Compressed Air versus 

Steam and Electric Transmission of Power, i 

CHAPTER II 

Structure and Operation of Compressors: Straight-Line and Duplex. Com- 
pound Steam End; Stage Compressors; Direct- and Belt-Driven or 
Geared Compressors. Comparison of Types. Relation of Work Done 
in Air and Steam Cylinders. Proportions of Cylinders. Compressors 
Driven by Water Power and Electric Motors, 8 

CHAPTER III 

The Compression of Air. Outline of the Theory. AppHcation of the Theory 
to the Operation of Air Compressors. Derivation of the Principal 
Formulae Relating to Isothermal and Adiabatic Compression. Modes of 
Absorbing the Heat of Compression, 49 

CHAPTER IV 

Wet Compressors. Hydraulic-Plunger and Injection Compressors. Injection 

Apparatus. Quantity of Injection Water Required, 75 

CHAPTER V 

Dry Compressors. Construction of the Water- Jackets. Circulation of Cool- 
ing Water. Piston Clearance and its Effect on Volumetric Capacity. 
Dry versus Wet Compression. Effect of Moisture in the Air under Com- 
pression. Effect of Injected Water, 81 

CHAPTER VI 

Compound or Stage Compressors. Theory and Operation. Single- and 
Double-Acting Stage Compressors. Construction and Functions of the 
Intercooler. Deductions from the Indicator Card of the Stage Com- 
pressor, -95 



Vlll CONTENTS 



CHAPTER VII 



Air Inlet Valves. Chief Requisites of. Poppet Inlet Valves: Their Construc- 
tion and Operation. "Skip-Valves" for the High-Pressure Cylinder of 
Stage Compressors. Ingersoll-Rand Hurricane-Inlet Valve. Johnson 
Valve. Humboldt Rubber Ring Valve. Leyner Flat Annular Valve. 
Arrangements for Admitting Inlet Air to the Compressor, 115 

CHAPTER VIII 

Discharge or Delivery Valves. Spring-Controlled Poppet Valves. Cataract- 
Controlled Poppets. Riedler Discharge Valve. Discharge Area for Air 
Cylinders, 136 

CHAPTER IX 

Mechanically Controlled Air Valves and Valve Motions. Mechanical Control 
for Discharge Valves. Norwalk, Nordberg, Laidlav^-Dunn-Gordon, Allis- 
Chalmers, Sullivan, Riedler, and Other Valve Motions. Cam-controlled 
Inlet Valve. Sturgeon Inlet Valve. Piston Valves, ....... 142 



CHAPTER X 

Performance of Air Compressors. Standards of Rating. Calculation of 
Horse-Power of Single-Cylinder and Stage Compressors. Mean Cylinder 
Pressure. Temperature of Compression- Elements of Air Indicator 
Card. Compressor Tests. A Record of Field Tests, w^ith Diagrams 
and Tables of Horse-Power and Costs. Summary, 159 



CHAPTER XI 

Air Receivers. Construction and Functions. Underground Receivers. Value 

of Cooling in the Receiver. "Receiver After-Coolers," 190 



CHAPTER XII 

Speed and Pressure Regulators for Compressors. Speed Governors. Air 
Cylinder Unloaders. Modes of Regulation for Steam- and Belt-Driven 
Compressors. Constant-Speed, Variable Delivery Gear, 196 



CHAPTER XIII 

Air Compression at Altitudes above Sea-Level. Consequent Reduction of 
Volumetric Capacity of Compressor. Relation between Compressor Out- 
put and Barometric Pressure. Mechanically Controlled Inlet Valves for 
High Altitudes. Stage Compression at High Altitudes, 216 

CHAPTER XIV 

Explosions in Compressors and Receivers. Discussion of Causes. Heat of 
Compression. Cylinder Temperatures and Flashing-Points of Lubricat- 
ing Oils. Examples of Explosions. Effect of Leakage of Delivery Valves. 
Precautions ior Preventing Explosions, 223 



CONTENTS IX 



CHAPTER XV 

Air Compression by the Direct Action of Falling Water. Theory of. Taylor 
Hydraulic Compressor. Descriptions of Plants at Magog, Province of 
Quebec, Ainsworth, B. C, Victoria Copper Mine, Mich., and Clausthal, 
Germany. Results of Tests, 235 



PART SECOND 

TRANSMISSION AND USE OF COMPRESSED AIR 

CHAPTER XVI 

Conveyance of Compressed Air in Pipes. Loss of Power. Loss of Pressure or 
Head. Discharge Capacity of Piping. D'Arcy's Formula. Richards's 
Formula for Loss of Pressure. Comparison of Results of Current 
Formulas. Compressed-Air Piping. Effect of Bends in Pipe-Lines, . 248 

CHAPTER XVII 

Compressed-Air Engines. General Considerations. Working at Full Pressure 
or with Partial or Complete Expansion. Ratios of Pressures and Tempera- 
tures due to Expansion in a Motor Cylinder. Corrections for Piston Clear- 
ance, etc. Nominal and Actual Cut-off. Work Done in a Motor Cylinder. 
Volume of Free Air Required. Cummings "Two-Pipe" System, . . 261 

CHAPTER XVIII 

Freezing of Moisture Deposited from Compressed Air. Causes and Preven- 
tion of Freezing. Influence upon Freezing of High Pressures in Trans- 
mission. Deposition of Moisture by Reduction of Pressure. Protection 
of Air Piping, . 274 

CHAPTER XIX 

Reheating Compressed Air. Appliances for, and Results of Reheating. Tem- 
peratures Employed and Consumption of Fuel. Construction and Opera- 
tion of Reheaters. Use of Reheaters for Underground Work. Wet and 
Dry Reheating, 279 

CHAPTER XX 

Compressed-Air Rock-Erills. General Description. Modes of Mounting 
Drills. Classification. Reciprocating Drills. Detailed Description of 
"Sergeant," Sullivan, Jeffrey, IngersoU-Rand "Arc-Valve," Murphy 
" Champion," Climax Imperial, Holman, Triumph, and Temple-IngersoU 
" Electric-Air " Drills. General Considerations as to Efficiency of Machine 
Drills. Consumption of Air: Normal and at Altitudes above Sea-Level. 
Factors Affecting Air Consumption: Examples from Practice. Proper Air 
Pressure for Machine Drills. Drill Repairs. Records of Work and 
Duty of Machine Drills. Conclusions, 294 



CONTENTS 



CHAPTER XXI 

Compressed-Air Hammer Drills. General Construction. Detailed Descrip- 
tions of Leyner, Hardsocg, Murphy, Sullivan, Ingersoll-Rand "Imperial" 
and "Crown," Waugh, and Stephens' "Climax Imperial." List of 
Makers. Depth of Hole and Speed of DrilHng. Records and Field of 
Work. General Conclusions, 343 

CHAPTER XXII 

Coal Cutting Machines. Classification: Endless Chain, Rotary-Bar, Disc or 
Circular Saw, and Reciprocating or Pick Machines. Detailed Descriptions 
of Jeffrey Endless Chain, Rotary-Bar, and Disc Cutters. Other Makes. 
Harrison, "New Ingersoll," Sullivan, Ingersoll-Rand "Radial-Axe," and 
Pneumelectric Pick Machines. Stanley Heading Machine. Auger 
Drills for Coal. Comparison of Coal Cutters, 380 

CHAPTER XXIII 

Channeling Machines. Applications and General Construction. Classifica- 
tion. Descriptions of Typical Machines. Depth of Cut and Speed of 
Work. Sizes, Specifications, and Weights of Sullivan and Ingersoll-Sergeant 
Channelers. " Electric- Air " Channeler, 409 

CHAPTER XXIV 

Operation of Mine Pumps by Compressed Air. Disadvantages of Using 
Ordinary Steam Pumps. Simple, Direct-Acting Pumps. Cylinder Dimen- 
sions of Simple Pumps. Volume of Air for Non-Expansive Working. 
Horse-Power. Regulation of Initial Air Pressure. Prevention of Freez- 
ing of Moisture. Compressed-Air-D riven Compound Pumps: Discussion 
of Modes of Using the Air. Application of Reheating, 423 

CHAPTER XXV 

Pumping by the Direct Action of Compressed Air. Pneumatic-Displacement 
Pumps. Merrill, Latta-Martin, and Harris Displacement Pumps. Pohle 
Air-Lift pump: Theory and Operation. Tests on Air -Lift Pumps. Ap- 
plication for Pumping Slimes in South-African Mills. Lansell's Air-Lift 
for Pumping in Mine Shafts, 438 

CHAPTER XXVI 

Compressed-Air Haulage for Mines. Compressed Air versus Electric Locomo- 
tives. Construction and Operation of Compressed-Air Locomotives. 
Modes of Dealing with Low Cylinder Temperature. Calculation for Pipe- 
Line and Charging Stations. Charging Apparatus. Calculation of 
Motive Power. Compressors for Charging Pneumatic Locomotives., De- 
tailed Examples of Compressed-Air Haulage Plants, .456 



ILLUSTRATIONS 



PAGE 

Figs, i and 2. — Laidlaw-Dunn-Gordon Straight-Line Compressor. Plan and 

Elevation, 10, 11 

Fig. 3. — Ingersoll-Rand Straight-Line Compressor, Class A-i, 13 

Figs. 4 and 5. — Laidlaw-Dunn-Gordon Duplex Compressor. Plan and 

Elevation, 14, 15 

Fig. 6. — King-Riedler Compound Vertical Two-Stage Compressor, . . . . 16 

Fig. 7. — Ingersoll-Rand Straight -Line, Two-Stage Compressor, 17 

Fig. 8. — Norwalk Compound Straight -Line, Two-Stage Compressor. Longi- 
tudinal Section, 19 

Fig. 9. — ^Norwalk Straight-Line, Two-Stage Compressor, with Simple Steam 

End, 20 

Figs. 10 and 11. — Leyner Straight-Line, Two-Stage Compressor. Plan and 

Elevation, ... 21 

Fig. 12. — Sullivan Straight -Line, Two-Stage Compressor. Longitudinal 

Section, Inset page 22 

Fig. 13. — Sullivan Corliss, Tandem-Compound, Two-Stage, Straight-Line 

Compressor, Class W C, 22 

Fig. 14. — Sullivan Duplex, Two-Stage Compressor. Longitudinal Section 

through Low-Pressure Cylinder, 23 

Fig. 15. — Ingersoll-Rand Cross-Compound, Two-Stage Compressor, Class O, 25 
Fig. 16. — Leyner Duplex, Two-Stage Compressor, with Simple Steam 

Cylinders, 26 

Figs. 17 and 18. — Riedler Cross-Compound Two-Stage Compressor; 15" and 

24'' X 36" Air Cylinders. Plan and Elevation, . . 27, 28 

Figs. 19 and 20. — AUis-Chalmers Cross-Compound Corliss, Two-Stage Com- 
pressor. Plan and Elevation, 29, 30 

Fig. 21. — Ingersoll-Rand Cross-Compound, Two-Stage Compressor, Class O, 31 
Figs. 22 and 23. — Laidlaw-Dunn-Gordon Duplex, Cross-Compound Compres- 
sor, with Two-Stage Air Cylinders. Perspective View and General Plan, 

Elevations and Sections, Inset and page ^^ 

Fig. 24. — Combined Air and Steam Cards, 35 

Fig. 25. — Duplex, 16" x 30" Risdon Compressor, Driven by 16 ft. Water-wheel, 37 
Figs. 26 and 27. — Water-Driven Risdon Duplex Compressor. Plan and 

Elevation, 38, 39 

xi 



Xll ILLUSTRATIONS 

PAGE 

Fig. 28. — Ingcrsoll-Rand Water-Driven Compressor, 41 

Figs. 39 and 30. — Rix Water-Driven Compressor at North Star Gold Mine, 

California. Side and Front Elevations, 42, 43 

Fig. 31. — IngersoU-Rand, Duplex, Two-Stage, Belt-Driven Compressor, . . 45 
Fig. 32. — Ingersoll-Rand Straight-Line, Belt-Driven Compressor, .... 46 
Fig. ^^. — Ingersoll-Rand Compressor, Duplex, Direct-Connected, Electrically- 
Driven; "Imperial" Type 10, Inset page 47 

Fig. 34. — Air Compression-Temperature Diagram, , . 54 

Fig. 35. — Reference Diagram, Elements of Air-Indicator Card, 57 

Fig. 36. — Air Indicator Card, 61 

Figs. 37, 38 and 39. — Air Indicator Cards, Showing Effect of Cooling, . . . 63 
Fig. 40. — Reference Diagram, Two-Stage Compression, No Clearance, ... 64 

Fig. 41. — Single-Stage Compression Diagram, with Clearance, 67 

Fig. 42. — Two-Stage Diagram, with Proportionate Clearance, 70 

Fig. 43. — Two-Stage Diagram, with Disproportionate Clearance, .... 72 

Fig. 44. — Humboldt Wet Compressor, 76 

Fig. 45. — Hanarte Wet Compressor, 77 

Fig. 46. — Air Cylinder of Nordberg Compressor, 82 

Fig. 47. — Air Cylinder, Class E, Laidlaw-Dunn-Gordon Co., 83 

Fig. 48. — Air Card Showing Effect of Clearance, 87 

Fig. 49. — Diagram of Effect of Clearance on Capacity of Dry Compressor, . 88 
Fig. 50. — Section of Air CyHnder, Showing Method of Reducing Clearance, . 89 

Fig. 51. — Section of Piston, Johnson Compressor, 89 

Fig. 52. — Diagram of Norwaik Two-Stage Compressor, ' 100 

Fig. 53. — Horizontal Intercooler. Ingersoll-Rand Co., 105 

Fig. 54. — Intercooler. Sullivan Machinery Co., 106 

Fig. 55. — Leyner System of Intercooling, 109 

Fig. 56. — Vertical Intercooler. Ingersoll-Rand Co., iii 

Fig. 57. — Combined Air Card of Two-Stage Compressor, 112 

Fig. 58. — Norwaik Poppet Inlet Valve, 118 

Fig. 59. — Laidlaw-Dunn-Gordon Poppet Inlet Valve, 119 

Fig. 60. — Diagram of Effect of Valve-Spring Resistance on Volumetric 

Capacity of Compressors, 121 

Fig. 61. — Air Card Showing Effect of Valve Resistance, 122 

Fig. 62. — "Skip-Valve." Norwalklron Works Co., 124 

Fig. 63. — Cylinder of Hurricane-Inlet Compressor. Ingersoll-Rand Co., . . 125 

Fig. 64. — Hurrican Inlet Valves. Enlarged Section, .127 

Fig. 65 and 66. — Johnson Air Valves, 129 

Fig. 67. — Humboldt Rubber Ring Valves, 130 



IIXUSTRATIONS Xlii 

PAGE 

Fig. 68. — Leyner Compressor. Part Section, Showing Flat Annular Air 

Valves, 132 

Fig. 69. — Leyner Annular Inlet Valve, 133 

Fig. 70. — Laidlaw-Dunn-Gordon Poppet Discharge Valve, 137 

Fig. 71. — Norwalk Poppet Discharge Valve, 138 

Fig. 72. — "Express" Poppet Valve. Riedler Compressor, 139 

Fig. 73. — Valve Motion of Low-Pressure Air CyHnder. Norwalk 

Compressor, 145 

Fig. 74. — Section of Air Cylinder of Nordberg Compressor, 146 

Fig. 75. — Section of Air Cylinder. Laidlaw-Dunn-Gordon Co., 147 

Fig. 76. — "Cincinnati" Valve Gear. Laidlaw-Dunn-Gordon Compressor, . 148 

Fig. 77. — Standard Air-Valve Motion. Allis-Chalmers Co., 150 

Fig. 78. — Sullivan Air Cylinder, Showing Corliss Inlet Valves, 151 

Fig. 79. — Riedler Air-Valve Motion, 153 

Fig. 80. — Details of Riedler Inlet Valve, 154 

Fig. 81. — Details of Riedler Discharge Valve, 155 

Fig. 82. — Cam-Controlled Inlet Valve, 156 

Fig. 83. — Sturgeon Inlet Valve, 157 

Fig. 84. — Diagram. Elements of Air Indicator Card, 167 

Fig. 85. — Air Card Diagram, 168 

Fig. 86. — Combined Cards, Two-Stage Nordberg Compressor, . . . .172 
Fig. 87. — Combined Cards, "Imperial Type 10" Two-Stage Compressor, . 172 
Fig. 88.— Card from 30^'' x 24" L. P. Air CyHnder of Style "O," Ingersoll- 

Rand Compressor, 173 

Fig. 89. — Card from 18^" x 24" H. P. Air Cyhnder of Style "O," Ingersoll- 

Rand Compressor, 173 

Figs. 90, 91, and 92. — Curve Diagrams, Compressor Plant No. i, . 175, 176, 180 
Figs. 93, 94, and 95. — Curve Diagrams, Compressor Plant No. 2, . 181, 182, 183 
Figs. 96, 97, and 98. — Curve Diagrams, Compressor Plant No. 3, . 186, 187, 188 

Fig. 99. — Curve Diagram, Compressor Plant No. 4, 189 

Fig. 100. — Vertical Air Receiver, Norwalk Iron Works Co., 191 

Fig. ioi. — Horizontal Receiver-Aftercooler, Ingersoll-Rand Co., .... 192 

Fig. 102. — Clayton Governor and Pressure Regulator, , 197 

Fig. 103. — Norwalk Pressure Regulator, 199 

Fig. 104. — Norwalk Pressure Regulator, 200 

Fig. 105. — Clayton Pressure Regulator, 201 

Fig. 106. — Rand Imperial Unloader, Sectional View, 203 

Fig. 107. — Sullivan Governor and Unloader, 205 

Fig. 108. — Ingersoll-Sergeant Regulator and Unloader, 206 



XIV ILLUSTRATIONS 

PAGE 

Fig. 109. — Laidlaw-Dunn-Gordon Air Governor, 208 

Figs, no, in, and 112. — Nordberg Constant-Speed, Variable -Delivery Com- 
pressor, Valve-Motion, and Regulating Gear, 210,211,212 

Figs. 113, 114, and 115. — Indicator Cards, Nordberg Constant-Speed, Vari- 
able Delivery Compressor, 214 

Fig. 116. — Air Cards Showing Results of Compression at Altitudes above 

Sea-Level, 217 

Fig. 117. — Taylor Hydraulic Air Compressor, 237 

Fig. 118. — Taylor Hydraulic Air Compressor, Detail of Head-piece, . . . 238 
Figs. 119 and 120. — Hydraulic Air-Compressing Plant at Kootenay, British 

Columbia, Inset and page 242 

Fig. 121. — Hydraulic Air Compressor at Clausthal, Germany, , , Inset page 246 

Fig. 122, — Expansion Curves of Steam and Air, 263 

Fig. 123. — Card Showing Work Done in Motor Cylinder, 267 

Fig. 124. — Leyner Compressed-Air Reheater, 285 

Fig. 125, — Cast-Iron Coils, Leyner Reheater, 286 

Fig. 126. — Sergeant Reheater, 287 

Fig. 127. — Rand Reheater, 288 

Fig. 128. — Sullivan Reheater, 289 

Fig. 129. — Sullivan Tappet Drill, 297 

Fig. 130. — Double-Screw Column Mounting for Rock Drills, 298 

Fig. 131. — "Sergeant" Rock Drill, Ingersoll-Rand Co., 301 

Fig. 132. — Spool-Valve and Chest, Sergeant Rock Drill, 303 

Fig. 133. — Sullivan "Differential" Drill (for Steam), 305 

Fig. 134. — Sullivan "Differential" Drill (for Air), . . 306 

Fig. 135. — Sullivan Tappet Drill, ' 307 

Fig. 136. — Sullivan Tappet Drill, Section, 308 

Fig. 137. — Jeffrey "Badger" Drill, 310 

Fig. 138. — Jeffrey "Badger" Drill, Diagram of Valves and Ports, 311 

Fig. 139. — Ingersoll-Rand "Arc -Valve" Tappet Drill, 313 

Fig. 140. — Murphy "Little Champion" Drill, 314 

Fig. 141. — Climax "Imperial" Drill, Inset page 315 

Fig. 142. — Holman Spool-Valve Drill, 316 

Fig. 143. — Holman Tappet-Valve Drill, 318 

Fig. 144. — "Triumph" Drill, 320 

Fig. 145. — Temple-Ingersoll "Electric-Air" Drill, 322 

Fig. 146. — "Water" Leyner Drill, 346 

Fig. 147. — Rotation Device, "Water" Leyner Drill, 347 

Fig. 148. — Hardsocg Wonder Drill, D-Handle, 349 



ILLUSTRATIONS XV 

PAGE 

Fig. 149. — Hardsocg Air-Feed Sloping Drill, 350 

Fig. 150. — Murphy Hammer Drill, with D-handle, for Sinking, 351 

Fig. 151. — Murphy Air-Feed Hammer Drill, 353 

Figs. 152 and 153. — Sullivan Hammer Drill, for Sinking and Plug Holes, . . 355 

Fig. 154. — Sullivan Air-Feed Hammer Drill, 357 

Fig. 155. — Ingersoll-Rand "Imperial" Hammer Drill, Types MV-i and 

MV-2, 359 

Fig. 156. — Ingersoll-Rand "Imperial" Hammer Drill, Type MC-12, . . 361 
Fig. 157. — Ingersoll-Rand "Crown" Hammer Drill, with Air-Feed, Type 

HC, 362 

Fig. 158. — Ingersoll-Rand "Crown" Hammer Drill, with D-Handle, Type 

HA, 364 

Figs. 159 and 160. — Ingersoll-Rand "Crown" Hammer Drill, Types HB 

and HC. Diagram of Valve Motion, 365 

Fig. 161. — Waugh Air-Feed Hammer Drill, 367 

Fig. 162. — Stephens's "Climax Imperial" Hammer Drill, . . . Inset page 369 

Fig. 163. — Sullivan Coal Pick, Working in a Thin Vein, 381 

Fig. 164. — Jeffrey Chain Machine, 382 

Fig. 165. — Jeffrey Chain Machine, Plan and Elevation, 383 

Fig. 166. — Jeffrey Chain Machine, Enlarged Plan and Elevation of Air 

Engines and Accessories, Inset page 384 

Figs. 167 and 168. — Jeffrey Disc Coal Cutter, Style 2 2-C, 386,387 

Fig. 169. — Jeffrey Disc Coal Cutter, Plan and Elevation, .... Inset page 388 

Fig. 170. — Harrison Pick Machine at Work, 390 

Fig. 171. — Sullivan Pick Machine, Mounted for Work, 391 

Fig. 172. — Harrison Pick Machine, Section, 392 

Figs. 173 and 174. — Rotary Engine for Operating Harrison Coal Pick, 

Top and Bottom Views, . . ^ 393 

Fig. 175. — Ingersoll-Rand Coal Pick, 394 

Fig. 176. — Ingersoll-Rand Coal Pick, Section, 395 

Fig. 177. — Ingersoll-Rand Coal Pick, Diagram of Valves and Ports, . . . .396 

Fig. 178. — Sullivan Coal Pick, Section, 399 

Figs. 179 and 180. — Ingersoll-Rand " Radialaxe " Coal Cutter, . . . . 400,401 

Figs. 181 and 182. — Pneumelectric Coal Puncher, 403 

Fig. 183. — Diagram of Gearing, Pneumelectric Coal Puncher, 404 

-Stanley Heading Machine for Collieries, 405 

-Sullivan Track Channeler, 410 

-Ingersoll-Rand Ram Track Channeler, for Marble, 411 

-Sullivan Rigid Back, Steam-Driven Channeler, . . , . . .412 
-Sullivan Adjustable Back, Air-Driven Channeler, 413 



Fig. 


184. 


Fig. 


185. 


Fig. 


186. 


Fig. 


187. 


Fig. 


188. 



Xvi ILLUSTRATIONS 

PAGE 

Fig. 189. — Ingersoll-Rand Undercutting Track Channeler, Type HF-3, , .415 

Fig. 190. — Ingersoll-Rand "Broncho" Channeler, 416 

Fig. 191. — Gibson-Ingersoll "Electric-Air" Track Channeler, with Swing 

Back, 421 

Fig. 192. — Merrill Pneumatic Pump, 439 

Fig. 193. — Diagram of Pohle Air-Lift Pump, 444 

Fig. 194. — Foot-piece for Air-Lift Pump, for Raising Mill Taihngs and Slimes. 45 1 

Fig. 195. — Diagram of Lansell's Air-Lift Pump for Mine Shafts, 454 

Fig. 196. — H. K. Porter Four-Wheel, Single-Tank, Compressed-Air Mine 

Locomotive, . 460 

Fig. 197. — Small H. K. Porter Compressed-Air Locomotive, 461 

Fig. 198. — Baldwin Six-Wheel Compressed-Air Locomotive, ...... 462 

Fig. 199. — Baldwin Four-Wheel Compressed-Air Locomotive, 462 

Figs. 200, 201, and 202. — Plan, Elevations, and Sections of Baldwin Com- 
pressed-Air Locomotive, Inset and pages 464, 465 

Fig. 203. — Compressed-Air Locomotive Charging-Station, 471 

Fig. 204. — Norwalk Locomotive Charging Compressor, 475 

Fig. 205. — Air -End of Ingersoll-Rand Three-Stage Locomotive Charger, . .476 
Figs. 206 and 207. — Low- and High-Pressure Air-Ends of Ingersoll-Rand 

Four-Stage Compressor, 477 

Fig. 208. — Perspective View of Ingersoll-Rand Four-Stage Compressor, . .478 
Fig. 209. — E. A. Rix Compressed-Air Locomotive for Empire Mine, Grass 

Valley, California, 481 



COMPRESSED AIR PLANT 



Part First 
PRODUCTION OF COMPRESSED AIR 



CHAPTER I 

INTRODUCTION 

One of the most important applications of the transmission of 
power by compressed air is the driving of machine rock-drills; 
and to the necessity of providing for these drills a power medium 
suitable for use in mines and tunnels has been due, more than to 
any other cause, the development of the modern air compressor. 

The time which has elapsed since the beginnings of this branch 
of engineering is short. The first percussion rock-drill, operating 
independently of gravity, was invented in 1849 by J. J. Couch, of 
Philadelphia. Though used only experimentally, it embodied the 
principal mechanical features of the modern machine drills, which 
have had such a striking influence in mining and tunnelHng. 
Couch's machine, together with its immediate successors, such 
as the Fowle drill (1849-51) and the Cave (Paris, 1851), 
were steam-driven and therefore unsuitable for underground 
work. In 1852, the physicist Colladon proposed the use of com- 
pressed air for operating rock-drills, in connection with the driving 
of the Mont Cenis tunnel, in the western Alps. His idea was de- 
veloped by Sommeiller and others between 1852 and i860, and in 
i"86i-62 an air-compressor plant was first used successfully at that 

1 



2 ^ COMPRESSED AIR PLANT 

tunnel. It was driven by water power and furnished air for 
ventilation as well as for the drills. 

The transmission of power by compressed air thus dates from 
about the middle of the last century. It is hardly necessary to 
say that the early air compressors were crude in both design and 
construction. Sommeiller's first plant, though of large size and 
effectual in fulfilling its purpose, had some resemblance in principle 
to the old hydraulic ram, possessing no moving parts except the 
valves. Piston compressors, driven by steam engines, such as the 
Dubois-Franfois, and more or less similar fundamentally to some 
of the wet compressors still in use, soon made their appearance. 
Probably the first compressors built in the United States were 
those employed at the Hoosac tunnel, in western Massachusetts, 
in 1865-66. The Burleigh, Norwalk, Clayton, and Rand compres- 
sors were among the earlier makes in this country. 

But the Mont Cenis tunnel, about eight miles long and com- 
pleted in 187 1, the first connecting link through the Alps be- 
tween the .railway systems of France and Italy, was undoubtedly 
the field where were solved on a large scale the initial problems of 
compressed-air production and use; and to Sommeiller is due the 
honor of having laid the foundations of new practice, by which that 
great work was brought to a successful completion. From 1857 
to 1 86 1 the tunnel headings had been progressing slowly and in 
the face of great difficulties. Drilling was done by hand labor 
and blasting by black powder, the average advance for this period, 
in each of the two headings, being only about one and a half 
feet per day. At this rate, even granting that the work could 
have been finished at all by the means employed, over forty 
years would have been required to connect the headings and 
years more to complete the enlargement to full section. With 
machine drills, the speed of advance in each heading rose to four 
and three-quarters feet per twenty-four hours and later, when 
dynamite was introduced, to a little over six feet; this average 
being maintained for a period of six years. 

Machine drills did not make their way into mining to any extent 
for some years after their successful application to tunnel driving. 



INTRODUCTION 3 

It is difficult now to name the mining district in this country where 
they were first used, but their most important trial was probably 
in the Calumet and Hecla copper mine, Michigan. After strong 
and concerted opposition from the miners, the Rand drill was 
introduced there in 1878, and the value of machine drilling for 
hard ground was soon demonstrated by decreased costs of drifting 
and stoping and higher speeds of advance. 

Compressed air has now a wide application in various branches 
of mechanical engineering and the arts and manufactures. In this 
book it is intended to deal only with its production and uses in 
connection with mining and tunnelling operations. Its two rivals 
in these fields of work are steam and electricity, regarding which 
a few general considerations may here be mentioned. 

As compared with steam, compressed-air transmission of pow- 
er is especially valuable and convenient for three reasons : flrst^ its 
loss in transmission through pipes is relatively small; second, the 
troublesome question of the disposal of exhaust steam underground 
is avoided ; third, the exhausted air is of some assistance in ventilat- 
ing the working places of the mine. In large mines, where steam 
may be carried thousands of feet, down shafts and through lateral 
workings, for operating pumping engines, etc., the disadvantages 
attending its use become very apparent ; the amount of condensa- 
tion is serious, even when the piping is provided with good non- 
conducting covering, and the working efficiency falls to an abnor- 
mally small figure. Furthermore, aside from the heat produced 
by the use of steam, it is rarely feasible to employ efficient con- 
densers for underground engines other than pumps, on account of 
the difficulty of obtaining the necessary condensing water and the 
additional space required. If the exhaust be discharged into the 
mine workings, even though they are large and well ventilated 
and the volume of the exhaust steam comparatively small, the 
temperature and quantity of moisture in the air would be con- 
siderably increased. Deterioration of the timbering is thereby 
hastened, the roof and walls of the workings are softened and 
slacked off, especially in collieries, and the mine atmosphere is 
rendered oppressive and unwholesome. The presence of hot steam 



4 COMPRESSED AIR PLANT 

pipes in confined workings, or in the narrow compartments of 
shafts, is also objectionable. 

Although the loss from condensation in long steam lines may 
be diminished by covering the pipe with efficient non-conducting 
material, still, even with the best covering, the effective pressure at a 
distant underground engine is greatly reduced, and very uneconomi- 
cal working is the result. On conveying steam a distance of several 
thousand feet the pressure may be reduced to half the boiler press- 
ure, or even less. For example, in the case of a pump, or other 
engine, situated 2,000 feet from the boiler and using 200 cubic feet 
of steam per minute at a boiler pressure of 75 pounds, with a 
four-inch mineral-wool-covered pipe, the effective pressure at the 
engine would be only about 58 pounds; or, with a poor covering, 
like some of the asbestos lagging often used, it might easily be as 
low as 35 pounds. For compressed-air transmission, on the other 
hand, the reduction of pressure for the same volume of air, size of 
pipe, and initial pressure, would be 9.3 pounds, giving a terminal 
pressure of 65.7 pounds. However, as the speed of flow in pipes 
for economical transmission is greater for steam than for air, a 
comparison based solely on piping of the same diameter cannot 
justly be made. In the above example, if the diameter of the pipe 
were smaller the gain in reduced radiation would outweigh the in- 
creased f rictional loss, and the net loss would be diminished. Since 
the frictional loss varies inversely, and the loss from radiation 
directly, with the diameter, the size of the steam pipe can be so 
proportioned as to produce a minimum loss under the given con- 
ditions. With compressed air the case is different, since the ques- 
tion of radiation is eliminated. If the diameter of the pipe be 
increased to 5 inches the loss of pressure, or head required to over- 
come friction, is reduced to 2.8 pounds and increasing the distance 
to one mile it would be only 7.4 pounds. Furthermore, the in- 
creased cost of the larger air pipe would be offset by the expense 
of the non-conducting covering necessary for steam transmission. 

Thus, compressed air may be conveyed long distances with but 
small loss of pressure, and is readily distributed for application to a 
variety of underground uses, for which steam is not practicable. 



INTRODUCTION 5 

Compressed air is always ready to do its work, and, aside from 
leakage of transmission pipes, which is in large measure pre- 
ventable, suffers no loss nor diminution of power when not in 
actual use. For performing work intermittently, at a distance from 
its source, it is therefore particularly valuable, because the air press- 
ure is maintained nearly constant during intervals of work, with- 
out further expenditure of power. With steam transmission, on 
the contrary, power is continually dissipated by radiation, whether 
in use or net, and a steam engine, when stopped for any length of 
time, loses much of its normal working temperature and becomes 
a receptacle for water of condensation. 

Though in mining compressed air is employed mainly for oper- 
ating machine drills, other applications are found in the driving of 
underground hoists and pumps in confined workings. Mechanical 
coal cutters, for mining bituminous coal, are sometimes operated 
by compressed air, and the employment of compressed-air loco- 
motives in mines and extensive tunnelling operations furnishes an 
example of its capacity for storing power, in contradistinction to its 
function as a power transmitter. The introduction of compressed- 
air drills has facilitated the rapid driving of long railroad and 
mining tunnels, which otherwise would have been greatly delayed 
or completed only with extreme difficulty. Had compressed-air 
power, together with the high explosives, not been available, it may 
well be doubted whether the great tunnels through the Alps and 
elsewhere, and the numerous long mine tunnels driven in recent 
years in this country, would have been at all practicable. 

Without attempting to review in detail the comparative merits 
of electricity and compressed air, it may be pointed out that the ap- 
plication of electricity for transmitting power in mines has increased 
enormously in importance during the past twenty-five years. The 
peculiar requirements of mine service have been in nearly all cases 
successfully met by modifications and adaptations of standard forms 
of electric apparatus. Both means of power transmission possess 
characteristics which adapt them particularly for underground 
work. But, although by virtue of its numerous successful applica- 
tions, electricity has become a rival of compressed air in most 



COMPRESSED AIR PLANT 

branches of mine work, their spheres of usefulness are not identical 
and the field is broad enough for both. It is often stated that the 
first cost of an electric plant is lower than that of an equivalent com- 
pressed-air plant. A broad generalization, however, does not cover 
the case. There is actually but little difference between the costs 
of the power plants themselves, the advantage being generally 
with the compressor. Considering the question of the transmission 
of a given pov/er, the cost of the electric conductor line for short dis- 
tances is much less than that of compressed-air pipe; but the cost 
of the electric line increases as the square of the distance, while the 
cost of the pipe line increases only as the first power of the distance. 
Hence, a point is soon reached where compressed-air transmission 
becomes the cheaper. It is in the greater efficiency of generation 
that electric power has the advantage. 

In one direction only has electricity failed hitherto to meet every 
requirement. While compressed-air drills, though far from being 
economical considered simply as machines, nevertheless admirably 
fulfil their purpose, no perfectly satisfactory electric rock-drill has 
yet been produced. However, this problem has for years been 
receiving much attention from electricians, both in this country and 
abroad, and there is reason to anticipate its successful solution in 
the near future. The Temple " electric-air " drill, brought out some 
four years ago, and already well tested under a variety of conditions, 
may be referred to here as a remarkably efficient and ingenious 
machine, but it is not an electric drill in the ordinary meaning of 
the term. It is rather a combination of a compressed-air drill, 
operated by a small, electric-driven compressor which is mounted 
on a truck close to the drill itself. As there is no exhaust, the same 
air being used over and over, one of the incidental advantages of the 
ordinary air drill is missing, namely, that of assisting somewhat in 
ventilating the mine workings, in those places where ventilation is 
m.ost needed. Keeping this in mind, together with such minor 
uses of compressed air as the cleaning of drill holes preparatory to 
charging, and driving out the smoke of blasting from working 
places, it seems doubtful whether, for underground mining, electric 
drills of any kind can be expected to supersede entirely those oper- 



INTRODUCTION 7 

ated by compressed air. Given the necessity for a compressed-air 
plant for the rock-drills, as is the case in most metal mines, it may 
often be more advantageous to provide the additional compressor 
capacity required for driving underground pumps, hoists, and other 
machines as well, than to erect a separate and distinct plant for 
generating electricity. 

Because of the view usually taken of the lack of economy 
in the operation of compressed-air drills, it has been customary in 
the past to consider compressed air in general as a form of power 
respecting which the questions of convenience and expediency are 
more weighty than the attainment of a high degree of efficiency. 
In recent years, however, as the principles of air compression have 
become better understood, a substantial improvement has taken 
place, not only in the design of the compressors themselves, but also 
in the installation of pipe lines and in the operation of the machines 
using the compressed air. The consequences of overloading a 
compressor, and thereby driving it beyond its proper speed, are now 
comprehended by every intelligent master mechanic as being wholly 
different from those produced by overloading a steam engine. 
The results of leaks in air pipes, and of using air mains of too small 
a diameter, are also understood and avoided. Better practice pre- 
vails in the field, and in the production, transmission, and use of 
compressed air a much higher total efficiency is now realized than 
was formerly thought possible. 



CHAPTER II 

STRUCTURE AND OPERATION OF COMPRESSORS 

An air compressor consists essentially of a cylinder in which 
atmospheric air is compressed by a piston, the power for driving 
which may be derived from a steam engine, water-wheel, or electric 
motor. The air cylinder is almost invariably double-acting, and 
as such is provided with inlet and discharge or delivery valves in 
each cylinder head. On the forward stroke the air is compressed 
by the advancing piston, while the decrease in pressure, or, as it is 
commonly termed, the tendency to form a vacuum, behind the pis- 
ton causes the inlet valves to open under atmospheric pressure, thus 
allowing the outside air to flow into the cylinder. At each stroke 
a certain volume of compressed air is forced from the cylinder 
through the discharge valves, into a pipe leading to a large reservoir 
or receiver, whence the air enters the transmission pipe or main. 

Before considering the operation and various appurtenances of 
the air and steam cylinders, it will be well to examine the general 
mechanical structure of the compressor and the modes of applying 
the power. Probably no single classification of air compressors 
can be made sufficiently comprehensive to present intelligibly all of 
their salient features. In attempting a classification three widely 
different bases of comparison suggest themselves. Firsts several 
clear distinctions result from a consideration of the general struc- 
tural characteristics of air compressors regarded purely as engines ; 
second, they may be classed according to the mode of dealing with 
the heat necessarily produced during compression of the air; and 
third, the numerous and varying types of valves and valve-motions 
devised in modern practice for controlling the distribution of the 
air in the compressing cylinders constitute a basis for comparison 

8 



STRUCTURE AND OPERATION OF COMPRESSORS 9 

which, though not so simple as the others, is in some respects 
quite as useful and important. 

The first classification only will be given here, the others being 
taken up respectively in Chapters IV to VI and VII to IX. Air- 
brake and gas compressors, vacuum pumps and other special forms 
of air-compressing machinery are not included, as this book will 
not deal with compressors other than those which are applicable to 
mine service. 

Under the first classification and taking the steam-driven com- 
pressor as the type form, four subdivisions may be named : 

1. "Straight-line" Compressors. In these, which are made by 
all builders, the steam and air cylinders are set tandem on a common 
piston-rod. They are always provided with a pair of fly-wheels, 
one on each end of the crank-shaft, v^^hich are driven by outside 
connecting-rods from a cross-head between the steam and air 
cylinders. Their structural form is thus simple and compact to a 
marked degree. Figs, i and 2 illustrate a Laidlaw-Dunn-Gordon 
straight-line compressor, with Meyer valve gear, a section of the 
steam cylinder being shown in the plan and of the air cylinder in 
the elevation. Fig. 3 is a perspective view of an Ingersoll-Rand 
straight-line compressor. 

2. Duplex Compressors, (a) Two engines are placed side by 
side, each being complete in itself and consisting of tandem steam 
and air cylinders, with their cranks set at 90 degrees on a common 
fly-wheel shaft. Each side of the duplex is in effect a straight-line 
compressor. They are almost invariably horizontal, and the steam 
cylinders are always nearest the crank-shaft. Figs. 4 and 5 show 
a type of the duplex compressor, (b) One air and one steam cylin- 
der, side by side, with a common crank-shaft, may in one sense 
be classed as duplex, though its operation is entirely different 
from (a) in the disadvantageous distribution of the load. While 
obsolete in America, this form is occasionally adopted by some 
European builders for special purposes. It is not well balanced, 
occupies at least three times the floor space of a straight-line com- 
pressor of the same capacity, and requires more expensive founda- 
tions. 



lO 



COMPRESSED AIR PLANT 



3. Compressors with 
Compound Steam Ends. 

(a) Duplex, horizontal, 
cross-compound; a simple 
or single-stage air cylinder 
being set tandem to each 
steam cylinder. This form 
is now rarely used. The 
considerations leading to 
the compounding of the 
steam end make it desirable 
to adopt stage compression 
for the air end. (b) Verti- 
cal compound; the air cyl- 
inders being placed respec- 
tively above the high- and 
low-pressure steam cylin- 
ders. This also is a some- 
what unusual design. It is 
advantageous in saving 
floor space, though this con- 
sideration is rarely of con- 
sequence at mines. As an 
example, the King-Riedler 
compressor may be cited 
(Fig. 6).* Some very large 
plants of this type, up to a 
capacity of 8,000 cu. ft. of 
free air per minute, have 
been built for South Afri- 
can mines. Vertical com- 
pressors have the disadvan- 
tage of being complicated 
and difficult to maintain 
and repair. 

* From American Machinist, 
Oct. i6th, 1902, p. 1,475. 




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STRUCTURE AND OPERATION OF COMPRESSORS 



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12 COMPRESSED AIR PLANT 

4. Compound or Stage Compressors, in which the air cylinders 
themselves are compounded. The air end may be of the double-, 
triple-, or quadruple-stage type, according to the air pressure to be 
produced.* Stage compressors are now made by nearly all 
builders in the United States, and compose the most important class 
of air-compressors for general use. (a) Straight-line form, as in 
(i). These have two-stage air ends, some having compound steam 
ends also. Fig. 8 shows the longitudinal section of a Norwalk 
compressor with compound steam cylinders, and Figs. 7, 9, 10, 
II and 12, respectively Ingersoll-Rand, Norwalk, Leyner, and 
Sullivan compressors, with simple steam cylinders, (b) Duplex 
steam end, with two-stage air cylinders. A longitudinal section of 
a Sullivan compressor of this class is given in Fig. 14, and a per- 
spective view of a recent type of Leyner compressor in Fig. 16. 
(c) Duplex, cross-compound steam end, with two- to four-stage 
tandem air cylinders. These are designed for large plants only. 
The two-stage type is widely used. Those having air cylinders 
of more than two stages are for special high-pressure service, 
such as furnishing air for underground compressed-air locomo- 
tives. Figs. 15 and 21 are perspective views of two of the latest 
designs of the Ingersoll-Rand cross-compound, two-stage com- 
pressors, class O. Figs. 17 and 18 show the general plan and side 
elevation of a Riedler, and Figs. 19 and 20 similar views of an 
AUis-Chalmers Corliss compressor of this class; Fig. 22 is a per- 
spective view, and Fig. 23 a reproduction of a working drawing, in 
plan and elevations, of a Laidlaw-Dunn-Gordon compressor, 
which will further illustrate this type. 

As based on structural characteristics, compressors may also be 
classified as: (a) Direct-driven by steam- or water-power — the 
motor end being directly connected with the air cylinders. Among 
water motors the bucket or impulse wheels are best adapted to this 
service; (b) Belt-driven from independent motors: steam-engines, 
water-wheels, or electric motors. These are built by most of the 
American makers, and are in common use for mine and other 

* It may be noted that the Norwalk Iron Works Co. was the pioneer in the 
field of stage compression, having begun in 1 880-81 to build this type of com- 
pressor for ordinary service. 




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service. Chain-driven and direct-geared compressors are also 
occasionally employed, as noted hereafter. 

So-called "half-duplex" compressors are furnished when re- 




FiG. 6.— King-Riedler Compound Vertical Two-Stage Compressor. 




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1 8 COMPRESSED AIR PLANT 

quired. They consist of either the right- or left-hand half of a 
duplex compressor, an extended crank-shaft and out-board pillow- 
block being provided temporarily. An advantage of this form is 
that, if only a comparatively small quantity of air is needed for a 
time — as during the development of a mine or the sinking of a 
shaft — one-half of a duplex compressor may be installed at 
first, the second half being readily added v^hen required. The 
capacity is thus doubled at a moderate cost. 

Comparison of Types or Compressor 

The straight-line compressor is largely employed for rather small 
plants or for temporary service. It is compact, strong, and self- 
contained, the entire engine being carried by a single bed-frame 
and requiring a relatively inexpensive foundation. The floor 
space occupied is much less than for the duplex form. The air 
and steam cylinders are just far enough apart to allow the cross-head 
and guides to be placed between them. From the cross-head the 
fiy-wheels are driven by connecting-rods on each side. By using a 
pair of fly-wheels each is made smaller and lighter than if there were 
but one, and the moving parts are better balanced. While useful 
for moderate air pressures and fairly constant loads, and satis- 
factorily filling an important field of work, the straight-line com- 
pressor is not capable of operating with the steam economy de- 
sirable and even essential in plants of large capacity; nor is it 
self -regulating at much less than, say, forty per cent, of its full load. 
These compressors are usually made of capacities from the smallest 
up to 1,700 or 1,800 cu. ft. of free air per minute, the last-named 
sizes developing from 275 to 300 horse-power. Further details of 
the operation and distribution of load in these compressors are given 
on page 32. 

The duplex compressor is always preferable to the straight-line 
for large plants. It is better adapted to varying loads, arising 
from differences of air pressure, because the resistance is more 
uniformly distributed throughout the stroke. By reason of its 
quartering cranks it may be run at extremely slow speeds without 
stopping on a center; and it is self -regulating and capable of deal- 



STRUCTURE AND OPERATION OF COMPRESSORS 



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20 COMPRESSED AIR PLANT 

ing economically with a range of load down to considerably less 
than one-quarter or one-third of its normal. As a rule, the friction , 
loss (total horse-power consumed by friction of the engine) of the 
duplex compressor is no greater and is often less than that of a ■. 
straight-line of the same capacity. For large Corliss compressors, 
in good order, this loss may be put at not over five to seven per 
cent.* While these figures are sometimes equalled by the best 




Fig. 9 — Norwalk Straight-Line, Two-Stage Compressor, with Simple Steam End. 

straight-line compressors, it is safe to say that the loss in the latter 
is generally higher. 

Of late years the Corliss type of engine has come into general 
use for driving large duplex compressors, especially when com- 
pounded in both steam and air end, as its valve gear is well adapted 
for dealing with the variations of air pressure under which com- 
pressors are usually called on to work. By the majority of builders 
the Corliss valve gear is employed, at least for large plants, for the 
air as well as the steam cylinders. 

The foundation of the duplex compressor is necessarily more 

* In this connection, see an article by J. Parke Channing, in Mines and Miner- 
als, May, 1905, p. 475, containing the results of an efficiency test on a 300-H.-P. 
compound, two-stage Nordberg Corliss compressor, at the Burra-Burra mine of 
the Tennessee Copper Co. Its efficiency was found to be 78.1 per cent, total. 
The horse-power consumed by friction was only 5.2 per cent. 



STRUCTURE AND OPERATION OF COMPRESSORS 21 



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STRUCTURE AND OPERATION OF COMPRESSORS 



23 




24 COMPRESSED AIR PLANT 

expensive than that of the straight-line, and must be substantially 
built if perfect alignment is to be maintained. Each pair of cylin- 
ders are solidly connected, either by trunk-frames or heavy tie- 
bolts. A complete girder-frame may be provided (Figs. lo, 14) to 
avoid any possibility of movement. The tandem steam and air 
cylinders on each side are best placed far enough apart to prevent 
the same portion of the piston-rod from passing alternately into 
each stuffmg-box. The reasons for this are: firsts the piston-rod 
is apt to wear differently in the two stufhng-boxes, so that it becomes 
difficult to keep them well packed and tight; second, in this con- 
struction the steam and air piston-rods are made in separate parts, 
coupled together between the cylinders. This is a matter of con- 
venience in making repairs, when it becomes necessary to take the 
compressor to pieces; also, the air valves, when of the poppet 
form and in the cylinder head, are more accessible. An incidental 
advantage of the duplex compressor is that, as each half is com- 
plete in itself, one side m^ay be disconnected for repairs or when a 
smaller capacity is temporarily desired. 

Compressors with Compound Steam Cylinders. The advantages 
in point of economy secured by compounding the steam end of air 
compressors are even more striking than in the case of ordinary 
stationary engines, for two reasons : First, because the conversion 
of power from one form to another is necessarily attended by some 
loss, and should therefore be conducted as economically as possible; 
second, because, as will be shown hereafter, the operation of 
compressing air involves particularly unfavorable load conditions. 
The valuable features of the duplex compressor become most 
apparent when the steam cylinders are compounded and furnished 
with a proper condenser. In plants of any size, a steam saving of, 
say, twenty per cent, may thus be readily attained, not only by getting 
the full expansive power out of the steam, but also by avoiding 
frequent loss of power due to imperfect speed regulation and con- 
sequent blowing off of air at the relief or safety valve. 

Stage Compressors in recent years have come into general use 
for mining and other service. It is now recognized that even for 
ordinary pressures of, say, seventy-five pounds, such as are com- 




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STRUCTURE AND OPERATION OF COMPRESSORS 



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STRUCTURE AND OPERATION OF COMPRESSORS 



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monly employed for machine drills, a saving in steam consumption 
can be realized. In elevated mountain regions, where so much min- 
ing is carried on, the advantages of stage compression are still greater 
than at sea -level, as is shown in Chapter XIII. The duplex form, 
with both steam and air ends compounded, exemplifies the highest 
type of compressor. There is no material increase in the number 
of moving parts, except valves; the greatest range of steam ex- 
pansion is obtainable, because the work done in the air cylinders 
is m.ore nearly equalized, and the compressor may be made self- 
regulating over its entire range of load. Thermodynamically, the 
efficiency of stage compression depends largely on the proper use 
of water-jackets for the cylinders, and the size and design of the 
intercooling apparatus between the air cylinders; a subject much 
better understood now than formerly. Stage compression is dis- 
cussed in detail in Chapter VI. 

Operation of Steam-driven Compressors. A steam-driven air 
compressor operates under peculiar conditions; appearing to 
work under a disadvantage which does not obtain in ordinary 
steam engines. This will be understood by inspecting the com- 
bined air and steam indicator cards of a simple straight-line com- 
pressor (Fig. 24). At the beginning of the stroke the air in front 
of the piston is at atmospheric pressure. As the piston advances 
the pressure at first increases slowly, while toward the end of the 
stroke it rises very rapidly. In other words, the resistance in the 
air cylinder varies from zero at the beginning of the stroke to its 
maximum near the end. The power developed in the steam cylin- 
der, on the contrary, when working as usual with a cut-off, is in 
exactly the reverse order. The initial steam pressure may be even 
lower than the final air pressure, though the mean effective pressure 
in the steam cylinder is greater than the mean effective in the air 
cylinder, as shown by the diagram. For example, with an initial 
steam pressure of sixty pounds, air may be compressed to eighty 
pounds or more. This result is obtained by the use of heavy fly- 
wheels and reciprocating parts, for carrying the engine over its cen- 
ters, storing up the surplus j)ower in the early part of the stroke, and 
giving it out toward the end. It follows that there is a marked want 





VATION 



wo-Stage Compressor 
ir cylinders, with det£ 




Fic. 21.— Laidlaw-Dunn-Gordon Duplex, Cross-Compound, Two-Stage Compressor. (End elevaUon of air end; 
longitudinal secUon through high-pressure steam and air cyUnders, with detail of discharge valve.) 




a 

o 

u 



34 COMPRESSED AIR PLANT 

of smoothness in the running of compressors, which causes severe 
strains in the moving parts. This is specially noticeable in the 
simple straight-line type, which, when the air in the receiver is up 
to gauge pressure, will often be brought almost to a standstill and 
barely turn over the centers. It would thus appear that only a 
small ratio of expansion in the steam cylinder could be employed, 
and in fact some of the older forms of straight-line compressors 
took steam throughout nearly the entire stroke. But the difficulty 
is met, and greater economy made possible, by the inertia of the 
fly-wheels. The dimensions of the steam and air cylinders in 
simple compressors are proportioned for a cut-off of from f to J 
stroke. 

In most of the simple straight-line compressors the steam cylin- 
der is provided with an adjustable cut-off valve (Fig. i ). This valve 
a is composed of two parts and, moving on top of the main valve^ 
controls ports in the latter through which steam is admitted to the 
main ports. It is operated by a separate eccentric on the fly- 
wheel shaft, and by means of the hand-wheel &, outside of the end 
of the valve chest, may readily be regulated without stopping the 
compressor, according to the varying pressure in the receiver. By 
manipulating this valve the compressor may be prevented from 
sticking on a dead center, notwithstanding considerable variations 
in receiver pressure. 

A number of arrangements have been devised in the past to 
equalize the power and resistance, by varying with respect to one 
another the positions of the air and steam cylinders and their 
cranks. For example, in the earlier forms of the Burleigh, De la 
Vergne, and Ingersoll compressors, the cylinders, instead of being 
parallel to each other, were placed at 90°, with the cranks at 30°. 
In the old Rand and Waring, of 1876, the cylinders were set at 45°,. 
the steam cylinder being of the oscillating pattern. The object of 
these and other similar devices was so to time the movements of 
the air and steam pistons that the power developed in the steam 
cylinder should be at its maximum when the air piston was just 
completing its stroke. But such constructions are deficient in 
strength and rigidity. They require heavier and more expensive 



STRUCTURE AND OPERATION OF COMPRESSORS 



35 



engine frames and foundations, and have not given satisfactory 
results. ~ 

In the duplex type, as already explained, the lack of equali- 
zation between power and resistance is minimized, the most favor- 
able distribution and the highest degree of economy being attained 
in duplex stage compressors with compound steam cylinders. 

Proportions of Cylinders. It is customary to build compressors 
with a short stroke, as this is conducive to economy in compres- 
sion, as well as the attainment of a proper rotative speed. A short 
stroke is of special importance in simple straight-line compressors, 
because the power and resistance are more nearly equalized than 




Fig. 24, — Combined Air and Steam Cards. 



with a long stroke. The motion is less jerky and there is less 
liability of stopping on a center. With a long stroke, and relatively 
small diameter of cylinder, the piston would travel some distance 
under a constantly increasing resistance; then, after the discharge 
valves open, it would advance a considerable distance farther under 
a uniform resistance, while adding nothing to the amount of useful 
work. It should be noted, however, that the loss of capacity of the 
compressor due to piston clearance is less for a long than a short 
cylinder of the same diameter. In ordinary single-stage, slide- 
valve compressors the usual ratio of length of stroke to diameter 
of steam cylinder is i| to i or if to i. In some makes, such as the 
older Rand compressors, the ratio was considerably greater, varying 
from I J or if to I. The length and diameter of steam cylinders in 
some recent designs are nearly equal. Quite different practice 



36 COMPRESSED AIR PLANT 

prevails, however, in the design of duplex Corliss compressors. In 
these are found such variations in the proportions of steam cylin- 
ders as: 12'' X 30'', 14'' X 42\ 20" X 42'', and 30'' X M'. 

The relative diameters of the air and steam cylinders depend 
obviously on the steam pressure carried and the air pressure to be 
produced. In mining operations there is usually but little varia- 
tion in these conditions. For rock-drill work, the air pressure is 
generally from sixty to eighty pounds. Of late, however, the appli- 
cations of compressed air for manufacturing purposes have so 
multiplied that some builders furnish compressors with steam and 
air cylinders of a great variety of proportions, for producing 
pressures of from ten to 120 pounds per square inch. 

Compressors Driven by Water-Power. When available, water- 
power furnishes a cheap and < convenient means of driving air 
compressors. Impulse or tangential wheels, such as the Pelton, 
Knight, orRisdon,are best adapted for this service, the wheel being 
mounted directly on the crank-shaft, as shown by Fig. 25. This 
cut is of a 16'' X 30'' compressor, built by the Risdon Iron Works 
for the Goleta Mining Co. It is driven by a sixteen-foot wheel under 
a head of 300 ft. Figs. 26 and 27 show plan and elevation of an- 
other compressor by the same makers. Plants similar to this are 
built by the Compressed Air Machinery Co., Ingersoll-Rand Co., 
and other makers. Since the power developed is uniform through- 
out the revolution of the wheel, water-driven compressors should be 
of the duplex type, in order to equalize the resistance as far as 
possible. The rim of the wheel is made extra heavy, to supply the 
place of a fly-wheel. This is illustrated by Fig. 28, of an Ingcr-i 
soil-Rand compressor driven by a Pelton wheel. 

To obtain the best efficiency, the peripheral velocity of an im- 
pulse wheel should be theoretically one-half the velocity of the jet 
of water from the nozzle. It follows that high heads of water in- 
volve correspondingly high peripheral velocities, and if the wheel 
be of small diameter a belt-drive would be required. But belting 
or gearing can generally be avoided, except when for any reason 
a turbine-wheel is adopted. Belt transmission is always disadvan- 
tageous, on account of the loss of power (say, eight to ten per cent.) 




O Xi 



0-, 



38 



COMPRESSED AIR PLANT 




STRUCTURE AND OPERATION OF COMPRESSORS 



39 




<U1 



40 COMPRESSED AIR PLANT 

and the cost of deterioration of the beking. Practically in all cases, 
an impulse wheel can be made of large enough diameter to run at 
a peripheral speed which will insure economical use of the water, 
while still giving a sufficiently low rotative speed for direct-con- 
nected compressor cylinders. In accomplishing this with very 
heavy heads, water-wheels are sometimes made of great size. 

Figs. 29 and 30 illustrate a well-known and interesting plant 
at the North Star Mine, Grass Valley, CaL, where a thirty-foot 
Pelton wheel drives a 300-horse-power, four-cylinder, two-stage 
compressor. The wheel makes sixty-five revolutions per minute 
under a head of 775 ft., with a single if inch nozzle. The cyl- 
inders are single-acting (to obtain more efficient cooling of the air 
and make it easier to detect piston leakage) and measure 30 and 18 J 
inches X 30 inch stroke. To avoid building excessively high 
foundations, as would otherwise be necessary for a wheel of this 
size, and also to furnish a substantial support for the gearing, the 
cylinders are set on angular frames at 30° to the horizontal. In the 
lower left-hand corner of Fig. 29, the intercooler is shown, sub- 
merged in the tail-race, a feature of the plant of no small advantage 
in producing a thorough cooling of the air without expense. The 
spur-wheel on the main shaft, with its accompanying pinion, is 
provided for operating the compressor, if necessary at any time, by 
a synchronous electric motor (shown in outline to the left, in Fig. 
30). At the extreme left of Fig. 30 is a small auxiliary water- 
wheel, for starting the motor and putting it in synchronism.* 

At the Morning Mine, near MuUan, Idaho, is another large 
water-driven two-stage compressor, of an entirely different de- 
sign. There are four cylinders, a high- and a low-pressure being 
set tandem on each side of a set of three Pelton wheels, mounted 
on the crank-shaft. A large volume of water, under a head of 140 
feet, is delivered through 8,000 feet of flume and 400 feet of pressure 
pipe, driving two 12-foot wheels. Two other streams, piped respec- 

* The bevel gears driven from the spur and pinion do not form part of this 
plant. They were used at one time to drive another compressor. For a more 
detailed description of this compressor, the illustrations of which were kindly sent 
to the writer by its designer, Mr. E. A. Rix, see American Machinist, November 
loth, 1898, p. 831. 




B 
o 
U 



G 

7 

00 

6 

H-l 



42 



COMPRESSED AIR PLANT 



lively I J and i mile, produce heads of 1,140 and 1,420 feet. These 
drive a 33-foot Pelton wheel (probably the largest in the world), 
placed on the compressor crank-shaft between the smaller wheels. 
The central wheel is driven by separate jets from the high-pressure 
lines, and on account of their difference in head, the diameter 
adopted for this wheel is a mean between the diameters theoretically 
necessary for obtaining a peripheral velocity properly proportioned 




Fig. 29.— Water-Driven Compressor at the North Star Gold Mine, California. 

(Side Elevation.) 



to each head. An actual mean peripheral speed of 8,000 feet per 
minute is attained. To control the water under these great heads, 
which correspond to pressures of about 490 and 610 pounds per 
square inch, slow-acting gate valves are provided, with by-passes 
for use in starting and stopping. The nozzles are arranged to be 
deflected clear of the wheel, in case it should be necessary to stop 
the compressor quickly. 

Each pair of cylinders are 33 J and 18 inches respectively X 42- 



STRUCTURE AND OPERATION OF COMPRESSORS 



43 



inch stroke, working at a piston speed of 560 feet. The low- 
pressure cyhnders compress to about 30 pounds, the high-pressure 
to 90 pounds. Inter- and after-coolers are placed in the tail-race 
of the smaller wheels. A positive valve-motion is employed for 
both inlet and discharge valves, which are of the Corliss type. On 
each side, parallel to the center line of the compressor and geared 
to the crank-shaft, is a long shaft. Geared to the latter in turn are 




Fig. 30. — Water-Driven Compressor at the North Star Gold Mine, California. 

(Front Elevation.) 



short shafts which carry the valve eccentrics. As the discharge 
valves must open when the pistons are moving at nearly their 
maximum velocity (800 feet per minute), an auxiliary dash-pot is 
provided for allowing them to open automatically under the 
cylinder pressure, the positive eccentric motion closing them. 

Indicator cards from_ this compressor show it to be highly effi- 
cient. An average of a number of cards gives mean pressures of : 
low-pressure cylinder, 17.86 pounds; high -pressure, 41.14 pounds; 



44 COMPRESSED AIR PLANT 

combined, 30.46 pounds. The mean theoretical adiabatic and 
isothermal pressures, corresponding to the combined mean are, 
respectively, 36.94 and 28.5 pounds. During the tests the ob- 
served temperatures were: coohng water, 38°; air at discharge 
from low-pressure cylinder, 135°; air at high-pressure inlet, 46°; 
high-pressure discharge, 140°; on leaving the after-cooler, 62°. 
Mean atmospheric temperature, 55° and of the cooling water 38°.* 

Provided there is a sufhcient volume of water, impulse wheels 
may be used with quite low heads, by introducing multiple nozzles, 
directed tangentially at two or more points of the periphery of the 
wheel. To prevent the water from splashing over the compressor, 
the wheel is enclosed in a tight wooden or iron casing. The force 
of the water may be regulated by an ordinary gate-valve; but if 
the head be great it is always desirable to use a special slow-moving 
gate (as noted above), to avoid danger of rupturing the pressure 
pipe in case the compressor is suddenly stopped. Turbines are 
obviously not so well adapted for operating compressors as the 
impulse wheels. A method of compressing air by the direct action 
of falling water is described in Chapter XV. 

Belt-Driven and Geared Compressors. These are often con- 
venient, and are furnished in a number of styles and sizes by com- 
pressor-builders. The fly-wheel is replaced by a large belt- wheel, 
with an extra heavy rim to give it sufficient weight. Fig. 31. The 
compressor illustrated is a recent design of the Ingersoll-Rand Co. 
In Fig. 32 is shown one of the older machines by the same makers. 
Power may be derived from an engine already installed for other 
purposes, or from a water-wheel or electric motor. Since electric 
transmission of power has come into general use in mining regions, 
a belt -drive from a motor is frequently advantageous when there 
is sufficient floor space. Some of the compressor-builders have in- 
troduced a "silent-chain" drive, for use when it is desired to place 
the motor close to the compressor and on the same bed-frame, and 
at the same time avoid the use of gearing. It has a high efficiency 
(about ninety-five per cent.) and may be employed for transmitting 
up to, say, 200 horse-power. 

* This plant, described in American Machinist, September 26th, 1901, was, \ 
like that at the North Star mine, designed by Mr. E. A. Rix. 




a 

o 
U 



m 



Q 



46 



COMPRESSED AIR PLANT 



Although a belt-drive is preferable to gearing, at least for a com- 
pressor erected on the surface, geared electric-driven sets are some- 
times used, a spur-gear on the crank-shaft engaging with a pinion 
on the armature. Single-reduction gearing v^ill generally answer. 
This design has been adopted even for large plants, as, for example, 
at a recent two-stage installation of the Compaiiia de Penoles, 
Mexico. By giving sufficient diameter and weight to the spur- 




FiG. 32. — Ingersoll-Rand Straight-Line, Belt-Driven Compressor. 

wheel, it not only produces the low piston speed necessary, but 
serves also as a fly-wheel. Raw-hide pinions are desirable to 
reduce the noise of the gearing. Induction motors are suitable for 
such service, as they are capable of running economically under 
wide variations of load. It may be added that the small, high- 
speed Christensen compressor is well adapted for gearing directly 
to a motor, thus forming a very compact machine for uses where 
lightness or portability is essential. Fig. 33 shows a longitudinal 
section through the low-pressure cylinder of a recent design of a 
direct-connected, electrically driven, duplex compressor, built by 



STRUCTURE AND OPERATION OF COMPRESSORS 47 

the IngersoU-Rand Co. One of the features of the compressor 
is that the frame is inclosed and all parts are self-oiling, except 
the piston and cylinder. The crank-pit contains the oil supply. 
The oil is picked up by the edge of the crank disc, and taken off 
at the top by a scraper. Part of it goes to a distributing tank, 
from which small pipes lead to the cross-head and guides. The 
main bearing and crank-pin are oiled direct from the scraper by 
a projecting trough on each side. For the bearing the oil is led 
to a channel in the cap, whence it passes through a series of holes 
drilled through the bearing liner. Some of this oil goes to the 
collar of the eccentric hub, from which it is carried to the face of 
the eccentric through two holes. As the eccentric is closed, the 
surplus oil is returned to the crank-pit. Thus the oil-feed is pro- 
portioned to the speed of the compressor, ceasing entirely when 
the compressor is stopped.* 

These direct-connected compressors are driven by direct-cur- 
rent induction or synchronous motors, the rotors of which are of 
large diameter, as shown, to produce a proper relation between 
the peripheral and rotative speeds. 

Under proper conditions an electric-driven compressor may 
be erected underground, near the point of application of the air 
power. Though some loss is inevitable in converting electric 
energy into compressed-air power, this may be offset in some cir- 
cumstances by considerations of convenience of installation. 
When the electricity is generated by water-power in large quan- 
tities, as in many Western mining districts, the cost per horse- 
power compares favorably with that of steam-driven compressors. 

Note. It is unnecessary, and in fact hardly practicable, in a 
book that is not intended to be a trade publication, to describe 
separately and in detail all the numerous makes of air compressors. 
In the foregoing chapter some of the wxll-known compressors are 
i instanced, for the purpose of illustrating various features of design. 
i The same remark may be made with regard to the descriptions of air 
valves, etc., in Chapters VII, VIII, and IX. It must not be under- 

* Similar self-oiling systems are applied to the compressors shown in Figs. 
I 15, 2 1, and 31. 



48 



COMPRESSED AIR PLANT 



Stood, however, that the compressors specifically referred to in this 
book are considered the only good ones, nor that the author, by 
omitting to mention and to insert cuts of all compressors, desires 
thereby to discriminate against those that are perhaps less well 
known only because they are the product of recently established 
builders. Most of the compressors made in Europe, including 
many excellent machines, are omitted altogether, though references 
to interesting features of the valve-motions of some of them will 
be found under the appropriate heads. 

An alphabetical list, which, while incomplete, comprises the 
names of most of the American compressor-builders, is given below. 



AUis-Chalmers Co. 

American Air Compressor Works. 

Chicago Pneumatic Tool Co. (Frank- 
lin compressor). 

Christensen Motor-Driven Compressor 
(Allis-Chalmers Co.). 

Clayton Air Com.pressor Works. 

Compressed Air Machinery Co. 

Franklin Iron Works. 

Heron & Bury Manufacturing Co. 



Ingersoll-Rand Co. 

Knowles Steam Pump Works. 

Laidlaw-Dunn-Gordon Co. 

Leyner, J. Geo., Manufacturing Co. 

McKiernan Drill Co. 

New York Air Compressor Co. 

Nordberg Manufacturing Co. 

Nor walk Iron Works Co. 

Rix Compressor and Drill Co. 

Sullivan Machinery Co. 



Vulcan Iron Works. 



I 



CHAPTER III 

THE COMPRESSION OF AIR 

In the production and use of compressed air occur serious losses, 
which to a large extent are unavoidable. Even in the best com- 
pressors the efficiency, or ratio of the force stored up in the com- 
pressed air to the work which has been expended in compressing 
it, rarely exceeds seventy-five per cent, and often falls below sixty 
per cent. To understand the causes of these losses it will be neces- 
sary^ to study the principles involved in the operation of compres- 
sing air. This study is advisable, also, before proceeding to a 
description of the air end of the compressor. Several definitions 
may first be given : 

"Free air" is a term commonly used in dealing with the subject 
of air compression. It is simply air at normal atmospheric press- 
ure, as taken into the cylinder of the compressor. But since 
atmospheric pressure varies with the altitude above sea-level, and 
with the barometric reading at any particular time or place, it 
follows that the expression ''free air" has no precise general sig- 
nification, with respect to the pressure, volume, and temperature of 
the air. At sea -level it is in reality " compressed air," at the normal 
atmospheric pressure of 14.7 pounds per square inch. As com- 
monly employed the term means air at sea -level pressure, and at 
a temperature of 60° Fah. 

The absolute pressure of air is measured from zero, and is equal 
to the assumed (or observed) atmospheric pressure plus gauge 
pressure; ordinary gauges registering pressures in pounds per 
square inch above atmospheric pressure. 

Absolute temperature is the temperature as measured from the 
''absolute zero" point, which is 491.4° F. below the freezing-point 
of water, or say 459° below zero Fahrenheit. For example, 60° F. 

49 



50 



COMPRESSED AIR PLANT 



of thermometric temperature is equivalent to an absolute tem- 
perature of 459° + 60° = 519° F. 

There are two fundamental laws governing the behavior of a 
perfect gas, when undergoing compression, which for all practical 
purposes are applicable also to atmospheric air. In discussing 
the problems of air compression, all the relations existing between 
volume, pressure, and temperature may be expressed in accordance 
with these laws. The first law (Boyle's) is: At constant tempera- 
ture the volume occupied by a given weight of air varies inversely 
as the pressure. This condition is expressed by the equation: 

P' V 



PV = P'V' 



constant; or — - 



V' 



in which 



V = the volume of the given weight of air (or gas) at the freezing- 
point and at a pressure P (F usually being taken as the volume in 
cubic feet occupied by one pound of air) ; F' = the volume of the 
same weight of air at the same temperature and at any pressure, 
P' (the pressures being absolute pressures). 

For example, to compress a quantity of atmospheric air at 
constant temperature to 0.147 of its original volume (the atmos- 
pheric pressure being 14.7 pounds), requires a pressure of 100 
pounds per square inch; when compressed to 0.074 of its original 
volume, the pressure required is 200 pounds, and so on. 

Table I \ 



Temperature 
Degrees Fah. 


Weight of 

one CiibicFoot 

in Pounds. 


Volume of one 
Pound in 
Cubic Feet. 


Temperature, 
Degrees Fah. 


Weight of 

one Cuoic Foot 

in Pounds. 


Volume of one 

Pound in 

Cubic Feet. 





.0863 


11.582 


no 


.0697 


14-345 


10 


.0845 


11.834 


120 


.0685 


14.596 


20 


.0827 


12.085 


130 


.0674 


14.847 


30 


.0811 


12.336 


140 


.0662 


15.098 


32 


.0807 


12.386 


150 


.0651 


15-350 


40 


.0794 


12.587 


160 


.0641 


15.601 


50 


.0779 


12.838 


170 


.0631 


15.852 


60 


.0764 


13.089 


180 


.0621 


16.103 


62 


.0761 


13. 141 


190 


.0612 


16.354 


70 


.0750 


13-340 


200 


.0602 


16.605 


80 


.0736 


13-592 


210 


-0593 


16.856 


90 


.0722 


13-843 


212 


.0591 


16.907 


100 


.0710 


14.094 









THE COMPRESSION OF AIR 5 1 

Table I * shows the weight and volume of dry air, at tem- 
peratures from o° to 212° F., and at atmospheric pressure. 

The production and use of compressed air, if governed solely 
by the law stated above, would be a simple matter. But during 
compression heat is generated, and when the air is allowed to re- 
expand to its original volume this heat is given up. Provided 
there is no transference of heat, the internal work, manifested by 
the increase of temperature, is independent of the time occupied 
by the compression. This condition is expressed by the second 
law, that of Charles and Gay-Lussac, viz.: When under constant 
pressure, the volume of a gas expands or contracts for each degree 
rise or fall of temperature, from freezing to boiling, by a constant 
fraction of the volume which it occupied at the freezing-point. 
Stated in another way, the volume of a gas under constant pres- 
sure is nearly proportional to the absolute temperature. The 
equation may be written: V = V (i + at^). The complete re- 
lations between pressure, volume, and temperature are expressed 
by the equation: P^"' = PV (i + al°), in which P' and V repre- 
sent the pressure and volume of a given weight of air (or gas) at 
t° Fah. above the freezing-point, P and V the pressure and volume 
of the same quantity of air at the freezing-point, and a the coeffi- 
cient of expansion of air, which is practically constant and is very 
nearly ^-^-y on the Fahrenheit scale. Hence, for a rise in tempera- 
ture of 1° F., the volume of the air increases by y^y of the volume 
occupied at the freezing-point, under the same pressure, 491° F. 
being the absolute temperature below freezing. 

The practical application of this law is that the development 
of heat reacts upon the air under compression, and increases the 
pressure due merely to the reduction in volume. By cooling the 
compressed air to its original temperature the pressure would be 
reduced to the normal amount, according to the first law. That is, 
the heat produced by the compression of a given volume of air 
corresponds in degree to the cold resulting from the re-expansion 
of the same quantity of air to its original volume and pressure. 
It is evident that this property of air has an important application 
in the production and use of compressed air. 

* From D. K. Clark and Appleton's "Applied Mechanics." 



52 COMPRESSED AIR PLANT 

Two Other statements may be deduced from what precedes: 
I. Under constant pressure the volume of air varies directly as the 
absolute temperature; 2. The volume being constant, the abso- 
lute pressure varies directly as the absolute temperature. 

The first of these statements is expressed thus: 

V V^ 

— = -r = constant, 
t f 

in which / and t^ are absolute temperatures; whence, from Boyle's 
law: 

V V^ 

P — = P'— = constant. 
t f 

For convenience, this constant is commonly denoted by R, and 

PV 

the general equation is written PV = R /, or — = R. 

i 

The value of R is found as follows : Since for a given weight of 

gas or air the density, D, is inversely proportionate to the volume, 

V = — ■, .08073 being the weight in pounds of one cubic foot 

.08073 

of dry air, at sea-level pressure (14.7 lbs.) and 32° F. The normal 
atmospheric pressure per square foot = 14.7 X 144 = 2,116.8 
lbs. If, therefore, one cubic foot be expanded by the application 
of heat to a volume of two cubic feet, the work done against at- 
mospheric pressure, per pound of air, will be ^ — -— = 26,220 

.08073 

foot-pounds. To double the volume, according to Boyle's law, 
would require the expenditure of 491.4° F. of heat; whence, in 
raising the temperature 1° F., the external work done by ex- 
pansion is : 

PV 26220 
/ 4914 

The heat generated during compression and corresponding to 
different pressures is shown in Table II, the volume at normal 
atmospheric pressure being i, at a temperature of 6o°Fah. 

From this table it is seen that the I'ate of increase of temperature 
is not uniform, but diminishes as the pressure rises. Thus, from 
I to 2 atmospheres the increase is 115.8°; from 2 to 3, 79.3°; from 



THE COMPRESSION OF AIR 



53 



Table II 





Absolute Pressures, 


Volumes in 


Final 
Temperatures, 
Degrees Fah. 


Corresponding 
Increases of 
Temperature. 


Pressure in 
Atmospheres. 


Pounds per Square 

Inch above 

Vacuum. 


Cubic Feet, 

Adiabatic 

Compression. 


I.OO 


14.70 


1. 000 


60.0 


00.0 


1-25 


^^-37 


0.854 


94-8 


34-8 


1.50 


22.05 


0.750 


124.9 


64-9 


2.00 


29.40 


0.612 


175-8 


115. 8 


2.50 


36.70 


0.522 


218.3 


158.3 


3.00 


44.10 


0-459 


255-1 


195 -I 


3-50 


51.40 


0.411 


287.8 


227.8 


4.00 


58.80 


0-374 


317-4 


257-4 


5.00 


73-50 


0.319 


369-4 


309-4 


6.00 


88.20 


0.281 


414-5 


354-5 


7.00 


102.90 


0.252 


454-5 


394-5 


8.00 


1 1 7 . 60 


0.229 


490.6 


430.6 


9.00 


132.30 


0.211 


523-7 


463-4 


10.00 


147-00 


0.195 


554-0 


494.0 


15.00 


220.50 


0.147 


681.0 


621.0 



3 to 4 atmospheres, 62.3°, etc. The quantity of heat developed 
during compression may be calculated by the following formula : * 

— - — X ^iap. log. — , m which 



Q 



Q = quantity of heat in thermal units. 

R = constant = 96.037 (French unit) or 53.37 (English unit). 

t =- absolute final temperature in degrees, corresponding to 
V (centigrade scale for French and Fahrenheit for Enghsh units). 

J = value of one thermal unit = 1,400 foot-pounds (or 778 
foot-pounds if English units be used). 

V and V = volumes of air in cubic feet, at beginning and end 
of compression. 

As the rise in temperature due to compression is proportional 
to the ratio of the final absolute pressure to the initial absolute 
pressure, the quantity of heat generated during compression to 
any given pressure, and the consequent work done, is greater at 
high altitudes than at sea-level. 

The above conclusions are illustrated by the diagram, Fig. 

* Zahner, "Transmission of Power by Compressed Air," p, 109. 



54 



COMPRESSED AIR PLANT 



"Volumes 



34.* It is, in reality, two diagrams, combined to save space. 
First, beginning at the lower left-hand corner, and curving up- 
ward, are the adiabatic and isothermal compression lines. 

Their intersections 
with the horizontal 
and vertical lines 
give the volumes 
of the unit of air 
when subjected to 
any given pressure, 
by reading the fig- 
ures at the top, 
and right- or left- 
hand margin of 
the diagram. The 
initial volume is 
taken as i, and the 
spaces between the 
vertical lines are 
each one - tenth. 
The resulting vol- 
ume is independ- 
ent of the initial 
temperature of the 
air. The corre- 
sponding pressure 
may be read in 
terms of either 
gauge or atmos- 
pheric pressure. 
Second, beginning 
at the lower right- 
hand corner of the 
and rising toward the left, are the lines of tempera- 
assumed initial temperatures being 0°, 60° and ioc° F. 

from "Compressed Air Production," by W. L. Saunders, several slight 
having been made in the adiabatic and isothermal lines. 



21 

20 

i 
19 

18 






















294.0 
279.3 
264.5 
249.9 




\ 


\ 


\ 


















\ 


\ 


\ 


















\ 


\ 


















16 
15 
14 

.a 10 




\ 


\ 




\ 














220.S 
205.8 
191.1 
176,4 
161.7 
147.0 
132.3 
117.6 
102.9 
88.2 
73.5 
58.8 
.44.1 
29,4 
14.7 
00 






\ 


\ 


\ 


















\ 


\ 


\ 


















\ 


\ 


\ 


\ 
















\ 


ij 




t 


















la 


















\ 


V° 


V 


















*A^ 


\ 




i 
<5 














Y 


\ 


\ 




1 












\ 


\ 


\ 


'3/ 
V 


1 




6 

5 
4 










\ 


A 


\ 




0/ 














\ 


\/ 




/ 














\ 


A 








3 
2 

1 












/ 


'\ 




\ 












^ 




/ 


V 


x\ 




_^ 


^^ 




-■^ 


^ 








A 


\ 



Temperature,Pah. 

Fig. 34, 



diagram, 
ture, the 

* Taken 
corrections 



THE COMPRESSION OF AIR 55 

The temperature corresponding to any given pressure is read on 
the lower margin. It should be observed that these temperature 
curves are those of adiabatic compression. 

It follows from the results obtained above that if the tempera- 
ture of the air be allowed to rise during compression an increase of 
work ensues. 

Isothermal and Adiabatic Compression. In accordance with 
the laws already stated, air may be compressed in two ways : 

Isothermal Compression. — ^The temperature is kept constant 
during compression, the heat generated being abstracted as fast 
as it is developed. In this case the pressure of the air varies ac- 

P' V 
cording to the equation P V = P'V, or — = — , and such com- 
pression may be called isothermal; that is, the compression 
curve of an indicator diagram would be an isothermal curve. 

Adiabatic Compression. — ^The temperature may be allowed to 
rise unchecked during the period of compression, as it will 
when there is no transference of heat, either by radiation or 
cooling devices. The rise in temperature increases the pressure 
that would be due to reduction of volume only. In other 
words, the pressure rises faster than the volume diminishes, and 

P' . . V 

— is no longer equal to, but is greater than — . 

This relation is determined by a consideration of the specific 
heats of air at constant pressure and at constant volume. The 
specific heat of any gas or vapor at constant pressure, C^, is the 
quantity of heat (in terms of heat units) required to raise the 
temperature of one pound of the gas i° F., the pressure being un- 
changed. The specific heat at constant volume, C^, is similarly the 
quantity of heat required to raise the temperature of the gas i° F., 
the volume being unchanged. Regnault's experiments have shown 
that for air C^ = 0.2375 ^^^ ^v = 0.1689. Of these C^ is the 
greater, because external work is done during a change of tem- 
perature, if the pressure be constant and the air free to expand; 
while, under constant volume, no work is done upon external 



56 COMPRESSED AIR PLANT 

resistances. When, as in adiabatic compression, the heat gen- 
erated reacts on the air under compression and increases the 

P' . V . 

value of — , to maintain the equation, — must be increased by 

an amount equivalent to the external work performed. The spe- 
cific heats may be expressed in heat units, as above; or, by mul- 
tiplying them by the mechanical equivalent of a heat unit (778 
foot-pounds = J), they are given in terms of foot-pounds and 
are then denoted by K; that is, 

JC, = K^and JC, = K,. 

Since K^ = C^ X 778 = 184.77. and K, = C, X 77^ = i3i-4. 
the ratio of these quantities gives: 



K^ _ 184.77 



1.406. 



K^ 131.4 
This ratio is commonly denoted by n, and is the exponent of the 

V . P' . 

power to which ^, must be raised to make it equal to -^'^' It is 

evident that n may also be expressed simply as equal to the 
ratio of the specific heat at constant pressure to the specific 
heat at constant volume: 

C^ 0.1689 
The general equation for adiabatic compression is therefore: 

\n= 1.406 

P ~ W^ 

Work of Compressors without Clearance. Isothermal Com^ 
pression. The work done by a compressor without clearance, 
and using isothermal compression, is represented by the area 
under the compression curve (Fig. 35). Let A B be an 
isothermal curve, A D representing any volume V of free air, 
and B C the volume V^ to which this quantity of air is com- 
pressed; the corresponding absolute pressures being respectively 

* A statement of the proof of this deduction is unnecessary here; it is given in 
several books on Thermodynamics, for example, in Perry's work on the Steam 
Engine, p. ^7,7,. 



p/ / V \ '^ 

PV«==FV-or--^ (-,) 



THE COMPRESSION OF AIR 



57 



P and P'. The curve is an equilateral hyperbola, and the work, 
W, represented by the area ABCD = Wi+W2-W3 



in which W, = area under A B = I V d 



V 



W2 = area under B C = P^ V, representing the work 
of expelling the air from the compressing 
cylinder. 

W3 = area under D A = P V, representing the negative 
work done by atmospheric pressure on the 
suction or intake side of the compressing piston. 



Note; V=BCorEC, 
according as the com- 
pression is isothermal 
or adiabatic. 




35. — Reference Diagram. 

Since P V = P' V, W2 and W3 cancel, so that the algebraic 
sum of 



W = W, 



w. 



w. 



= fvd 



V 



w 

P'V 

To integrate this expression, substitute for P its equivalent -^rz- 

(from the general equation for isothermal compression), whence: 

r^ dv r^ dN 

Jy' V "^v V 



Integrating: W = P' V X Nap. log. (-777)* (2) 



* The Naperian or hyperbolic logarithm of a number, generally written " log.e,' 
is obtained by multiplying the common logarithm by the constant 2.302585. 



58 COMPRESSED AIR PLANT 

The equation may also be written: 

W = PVlog.,(|^), (3) 

which is a form convenient for use in making air compressor cal- 
culations. When expressed in foot-pounds (by putting V in terms 
of cubic feet, and P, P' in pounds per square inch), the equation 
takes the form: 

W= 144 PV log., (1^) (4) 

If V, the intake capacity of the cylinder in cubic feet, be de- 
notea by L, the equation becomes: 

W=i44PLlog., {—) (5) 

which is the general equation for the work of compressors oper- 
ating isothermally and without clearance. 

Work of Compressors without Clearance, Adiabatic Com- 
pression. The expression for the work done in compressing air 
adiabatically is deduced as follows, referring to Fig. 35. The 
line A D represents the initial volume, V, of air at normal at- 
mospheric pressure, to be compressed, and the line E C the final 
volume V^ occupied by the same quantity of air at the end of 
the stroke of the piston. In undergoing this change of volume, 
the pressure increases from P to P^ (absolute pressures), and the 
resulting compression line A E is an adiabatic curve, following 
the law: 

P V^ = P' V'" = C (constant), or P = ^^ (6) 

In terms of foot-pounds, the total work done in the compress- 
ing cylinder is: 

W = (W, + W, - W3) ■. (7) 

in which: 

Wi = the work of compression. 

W2 = work required to force the compressed air out of the 
cylinder, into the receiver. 



THE COMPRESSION OF AIR 59 

W3 = work done by atmospheric pressure on the suction side 
of the piston, while the inlet air is entering the cylinder. 

First. — The work W^ in foot-pounds, done during compres- 
sion, is represented by the integral expression: 

W, = f 144 P^V 

Substituting the value of P, now expressed in pounds per square 
inch, from equation (6) : 

r"^ CdY 
W. = Wv.^ ■■■ (8) 

Integrating between the limits V and V : 

pyd-n) _ y/(i-n)-, 
W. = I44CL ___]..... (9) 

Dividing the second member of the equation by (— i) and sub- 
stituting for C its value P V^ : 

^^, 144 P V- r I I -1 , , 

i44Pvrv(^ -| , , 

P' V V /P'\^ 
But, since — = rrr^, -777 = I ~^ ) " > which, raised to the n — i 

power, gives: 

yCn- I) / p/yn-i 

v^(^= \t) " ^''^ 

Substituting this value in (11) : 

w, = --M|f[©-^.] ,,, 

Second.— The work W2, of expelling the air from the cylinder, 
= i44P'V' (13a) 

V . P' /V\'' 

Multiplying by — both members of the expression, — = \-y}J : 



PV 



(^,) ; whence, FV^ = PV(— ). 



6o COMPRESSED AIR PLANT 

But, 



- - [-y) « and (-J = i p-j - ^^^^^ 

/ p/X n -I 

p' V'= PV ( — j w ; which, substituted in equation (13a), gives: 

W, = I44FV^ = I44PV(-|^)^' (14) 

Third. — The work W3, done by atmospheric pressure on the 
back of the piston, = 144 P V (15) 

Taking the algebraic sum of W^, W2 and Wg, from equa- 
tions (13), (14) and (15), and substituting in equation (7): 

W = 144 j^C©"^ - I ] + P V (1-')"^'- P V j ; 

whence, by reducing to a common denominator: 

Pv[QV_ i] + („-!) pv(^)V_(„_i)Pv 

W = 144 

n — 1 

and cancelling: 

w-^?[(fF--] <■«) 

which is the general expression for the work of single-stage com- 
pressors, with adiabatic compression, and when clearance is zero. 
The relations between the two conditions of compression are 
represented graphically by Fig. 36. By laying off to scale the 
volumes of air on the horizontal line of the diagram, the corre- 
sponding pressures at different points of the stroke of the compress- 
ing piston are measured on the verticals. The adiabatic curve rises 
more rapidly than the isothermal, according to the law. There- 
fore, in compressing adiabatically a quantity of air to a given 
volume, more work is expended than if the compression were 
effected isothermally. Perfect isothermal compression cannot be 
attained in practice. Even w^ith the best cooling arrangements the 




THE COMPRESSION OF AIR 6 1 

compressor would have to run at an extremely slow speed, and be 

of very large size, to approach closely the condition of isothermal 

compression. On the other hand, if the air 

compressed adiabatically could be kept hot 

until used, the loss of the additional work 

which was expended in compressing it 

would be prevented. But neither can this 

be done. The air 

is almost always ^a^^^^i^^^^^ 

conveyed to con- 

siderable distances 

before it is used, ^^^- ^^• 

and the loss of heat by radiation from the pipes soon reduces the 
pressure to that corresponding with the temperature of the sur- 
rounding atmosphere. In practice, therefore, neither of these 
theoretical methods of compression is possible; a combination or 
modification of the two is employed, the net result depending upon 
the degree of perfection of the compressing engine and of the cool- 
ing arrangements provided. 

As shown by Fig 36, the actual line of compression must 
lie somewhere between the adiabatic and isothermal lines 
provided there is no leakage past the piston. When com- 
pressing in a single cylinder to sixty or eighty pounds' pressure, 
and at a piston speed not exceeding 300 feet per minute, it is 
probable that about one-half of the total possible cooling is all that 
may be expected.* The aim is to bsgin compression with the air 
at a low initial temperature, and to bring the compression line as 
close as possible to the isothermal line. Next, it is of the utmost 
importance that the air shall be cooled thoroughly during com- 
pression and before it leaves the cylinder. Any subsequent cooling, 
whether in the receiver or in the air main, must entail loss. 

As a matter of fact, the abstraction of heat during compression 
in ordinary practice is very imperfect. Some distance must be trav- 
ersed by the piston, in compressing the air, before there is any 
considerable rise in temperature, and until the temperature does 

* Frank Richards, " Compressed Air," p. 66. 



62 COMPRESSED AIR PLANT 

rise no cooling can be effected. In other words , the abstraction of 
heat does not begin at the beginning of the stroke. The temper- 
atures of the air taken into the cylinder and of the water used for 
cooling are likely to be nearly the same, so that all the possible 
reduction of temperature in any one cylinderful of air must take 
place in a period of time less than that occupied in making the 
stroke. Most of the cooling is done necessarily in the latter half 
of the stroke. It should be noted, moreover, that soon after the 
compressor begins running the cylinder itself becomes quite hot 
and heats the air during intake. For this reason the total amount 
of cooling to be effected is greater than that which is required to 
abstract the heat developed during the compression of a given 
volume of air to a given tension. In modern dry compressors of 
fairly large size, and running at full working speed, the com- 
pression line is usually much nearer the adiabatic than the iso- 
thermal curve, and often follows the adiabatic curve quite closely. 
There are two methods of absorbing the heat produced by 
compression : 

1. By introducing cold water into the air cylinder. 

2. By cooling the cylinder from without, enveloping it in a 
cold-water jacket. 

Machines of the first class are known as ''wet compressors"; 
those of the second, "dry compressors." 

The values of the coefficient n in the equation already given, 
pr . Y . « 

— = f — ) , have been found for the different systems of compres- 
sion. As has been stated, in the case of purely adiabatic compres- 
sion, with no cooling arrangements, w= 1.406; in ordinary single- 
cylinder dry compressors, provided with a water-jacket, n is 
roughly 1.3, while in the best single-stage wet compressors (with 
spray injection) n becomes 1.2 to 1.25. In the poorest forms of 
compressor the value 71=1.4 is closely approached. It should be 
added that for large well-designed compressors with compound air 
cylinders and efficient intercooling, the exponent w, referred to the 
combined indicator cards, may be as small as 1.15. This result 
has been obtained, for example, from a 2,00c horse-power, two- 
stage compressor at Quai de la Gare, Paris. 



THE COMPRESSION OF AIR 



63 



The diagrams, Figs. 37, ^8, and 39, show the relative positions 
of the several compression lines, the areas between the compression 
and isothermal lines being shaded in each case. These are not 
actual indicator diagrams. They are intended approximately to 
represent the relations between the different lines, under the con- 
ditions named. 



Compression 
"Without Cooling 




Fig. 37. 



Effect of 
Water Jacket 




Fig. 38. 



Effect of 
Spray Injection 




Fig. 39. 

Work of Two-stage Compressors, without Clearance.*— In 

this system of compression the air is brought up to a certain 
pressure in one cylinder; passes thence to an intercoolirg 
chamber, or intermediate receiver, in which the temperature of 

* For the construction and operation of stage compressors see Chapter VI. 



64 



COMPRESSED AIR PLANT 



the air is reduced, and finally enters a second cylinder, where 
the compression is carried to the desired terminal pressure. It 
is customary to design the compressor with a cylinder ratio that 
will divide the work equally between the cylinders, but changes 
of conditions of operation other than those contemplated may 
destroy this equahty. In the following discussion,* it is as- 
sumed that the same quantity of work is done in each stage. 

An inspection of the diagram, Fig. 40, shows that there must 
be some best intermediate receiver pressure, for which the total 



> 
F 


E Q 






\ \ \ 






\ \ 1 






^ \ * 






^ \ ^ 






^ \ \ 






^ \ \ 






^,V-^^ 7"^ =0 






1^\\ \ / 






\V \r./ 






VN V 






%A 






'^^x \ 






N-^"^ X 






^^^N. 


















^ X. 






>» >s^ 












^^^^\^ 






"^^J*^ B 




y 



Fig. 40. — Reference Diagram, Two-Stage Compressor, with no Clearance and 
Perfect Intercooling. 



work of compression will be a minimum. For, if this receiver 
pressure approach either P or P'' (corresponding respectively to 
the points B and G on the compression curve), then would the 
compression approach single-stage work and the entire compres- 
sion line would lie along B C G. But, with the intercooling re- 
ceiver at any intermediate point, the broken line B C D E is fol- 

* I desire here to acknowledge the valuable assistance of Dr. Charles E. Lucke, 
Professor of Mechanical Engineering in Columbia University, who kindly gave me 
the use of his original notes on the method of analysis employed in this discussion 
of the theory of stage compression and the effect of clearance in the different work 
cycles of adiabatic compression. 



THE COMPRESSION OF AIR 65 

lowed, the saving in work over single-stage compression being 
represented by the area C D E G. 

The net work of the compressor, W, represented by the area 
A B C D E F, is equal to the work of area A B C H of the first 
stage plus the work of area H D E F of the second stage, or W = 
Wi + W2. Let the condition of perfect intercooling be assumed; 
that is, the hot air discharged from the first cylinder is cooled 
in the intermediate receiver to the initial temperature of the in- 
take air. The work cycle in each cylinder is the same as 
that of single-stage adiabatic compression, as expressed by 
equation (i6), but with two additional symbols for pressures and 
volumes. 

initial volume of free air in first cylinder, 
initial volume of air in second cylinder, 
initial absolute pressure (atmospheric pressure), 
terminal absolute pressure in first cylinder, as- 
sumed to be also the intermediate receiver 
pressure and therefore the initial pressure in 
the second cylinder. 

O F = P'' = terminal absolute pressure in second cylinder. 



AB 


= V 


HD 


^ Y" 


OA 


= P 


OH 


= P' 



(17) 



XT T.r PVwr/F\^^ -1 

Hence: W^ = L V^J " ~ ^ \ ^^^ ^^^^ 

p'V"wr/p''\^iiL^ n 

W2 = L \ P^ / ^ — I . . . second stage 

But, assuming the intercooling to be perfect, P V = P' V, whence: 

Since the best receiver pressure, P', is that for which W is a mini- 
mum, by differentiating and placing the first differential coeffi- 

cient^p^=o: 



dV n 



p— (F) 



66 COMPRESSED AIR PLANT 



Whence 



(F) "^ _ {^")^-' 

n -1 2n - I 



or (P0~^ ~^^ = (P P") " fi*oni which P' = VP P", an expres- 
sion for the best receiver pressure. 

Dividing both terms by P : 

F (P X P'O* /^P"\* 



P P 



(f)' 



(P''\ i P'' V" 

-— ) = ' = -— -. Substituting these values in equa- 
P ^ \/VV" p' ^ ^ 

tion (17), remembering that V V = P V and expressing the 
work in foot-pounds: 

w = — 57^7^11^)^^ - ^J (^9) 

which is the equation for two-stage compressor work, in terms 
of the initial volume and initial and terminal pressures, with 
perfect intercooling and best receiver pressure. 

By a method similar to the above, the expression for the work 
of three-stage compression may also be deduced : 

3 



W 



i44]pv^r/p- V— 1 t X 



in which V" is the terminal pressure in the last, or high-pressure, 
cylinder. 

Effect of Clearance in the Compressing Cylinder. In the pre- 
ceding pages expressions are deduced for the work of compression 
with no allowance for the clearance volume of the cylinder. From 
a mechanical engineering and structural point of view, the ques- 
tion of piston clearance is taken up in Chapter VII. It is neces- 
sary here to discuss the cycles of operation of single-stage and 
two-stage compression with clearance. While the work done is 
the same as that shown by the general equations found for iso- 
thermal and adiabatic compression, the work per unit of cylindei 
volume, or displacement, will be changed, because of the re- 
expansion of the clearance air. In other words, clearance affects 



THE COMPRESSION OF AIR 



67 



the volumetric output of the compressor, but not the work of com- 
pression per unit of volume of air taken into the cylinder. 

Work of Compressors with Clearance. Isothermal Compres- 
sion. Fig. 41 is a general reference diagram, in which B C 
represents an isothermal curve. 




J F H G 

Fig. 4 1 . — Reference Diagram. Compressor Working Isothermally, with Clearance. 



LetEB 


= M F 


= P 










GC 


= F D 


= P' 










FE 


= M B 


= V 










CD 


= V 












JF 


= ND 


= clearance volume 






FH 


= (FE 


- H 


E) 


= MA = 


(M 


B 



A B) = V'', or 

volume occupied by the re-expanded clearance 
air. 

According to the diagram areas, the total 
Net Work A B C D = compression work and delivery work 

O B C N — re-expansion work 

ADN. 
For the work of a compressor without clearance, and isother- 
mal compression, the general expression, W = P V log.^ ( -5- ) , has 

already been deduced. This applies to the areas bounded by 
the two horizontal lines, the vertical line and the compression 
line. Similarly, the re-expansion work represented by the area 



68 COMPRESSED AIR PLANT 

FDAH (under tHe curve D A) = P V" log.^ (~) . 
Hence, W = P V log., {^) - P V" log., (|^) 

= P (V - V") log., (1^) (21) 

Replacing (V — V^) by L, which represents the intake capacity 
of the compressing cyHnder, neglecting heating during suction, 
and expressing P in pounds per square inch: 

p/ 

P 



W =144 PL log., (1^) (22) 



Comparing equations (5) and (22) it is seen that they are iden- 
tical, as noted above; but it must be remembered that the volume 
of air actually taken into the cylinder at each stroke and com- 
pressed is reduced by the clearance space, and hence the volu- 
metric capacity of the compressor is also reduced. Moreover, 
neither in this work cycle, nor in that for adiabatic compression, 
is any account taken of the heating and cooling effects which occur 
during intake and compression, nor of frictional and other losses 
which affect capacity and work per unit of air. These points 
are discussed elsewhere in this chapter and in Chapters IV, 
V, VI, VII and X. 

Single-stage Adiabatic Compression, with Clearance. The 
reference diagram, Fig. 41, may be used here also, by assuming 
the line B C to be an adiabatic curve. But, though the work 
areas are designated as under isothermal compression, and their 
significations are identical, their numerical values are different. 

From equation (16), the work corresponding to area 

OBCN = 

Wi = -^ \~w)~^ ~^ ' ^^^' similarly, the work corre- 
sponding to the area O A D N = 



THE COMPRESSION OF AIR 69 

^.,.n.e,W^^-^^l^^^P-%{^y-^'-q (.3) 

Replacing (V — V') by L (the intake capacity of the cyhnder, 
with clearance) : 

"^^-V^T IW) ^ - '] (^4) 

Since V in equation (16) may be replaced by L, a comparison 
of this equation with (24) shows that the work is the same per 
unit of volume of air admitted to the cylinder; but, the volumetric 
output is reduced by the clearance. 

Though the pressure-volume formulas serve for most purposes, 
it is sometimes convenient to have the work expressed in terms of 
cylinder displacement and volumetric efficiency: 
Referring to Fig. 41 : 

Let D = displacement volume of cylinder in cubic feet, or the 
effective area X stroke, represented on the 
diagram by M B = F E. 
C = clearance expressed as a fraction of D; whence C X 
D = V^ = clearance volume, represented by 
N D = J F, and D (i + C) = total cylinder 
volume in cubic feet, represented by J E. 
V"' = volume of re-expanded clearance air. 
E = intake free air capacity = F E — F H. 

E =- volumetric efficiency =-^ = the ratio of the length of 

the actual admission line, A B, to the total dis- 
tance swept through by the piston. 

Then: V = D (i + C), and V'' = V, (|^) - = C D (|^)^ 

Whence V - V''' = D [i + C - C (^')^] = L 
Substituting this value of V - V' in equation (23) : 



70 



COMPRESSED AIR PLANT 



which expresses the work in terms of displacement and clear- 
ance, with pressures in pounds per square inch. 

Since E = -^, L = ED=d[i + C- c(p-)^], as above; 

whence, substituting in (25) : 

W = ^"hb[©=.'-.] ,=., 

Work of Two-stage Compressor, with Clearance. (Fig. 42 .) 
By a method of deduction similar to the preceding, it- may be shown 




Fig. 42. — Reference Diagram. Work of Two-Stage Compressor, with Clearance. 



that the best intermediate receiver pressure for two-stage com- 
pression with clearance is the same as that for stage compression 
without clearance. This follows because the receiver pressure 
is a function of the compression line and not of the re-expan- 
sion line. It will be found, also, that, unlike the w^ork cycle 
for stage compression without clearance, there is here an 
unequal division of work between the two cylinders. (See 
also Chapter VI.) 

The general equation for two-stage compression, with propor- 



THE COMPRESSION OF AIR 7I 

tionate clearance in both cylinders and perfect intercooling at best 
receiver pressure, is * : 



The diagram (Fig. 42) shows a single, continuous re-expansion 
line, F A, which is taken to represent the re-expansion lines of 
both cylinders. This evidently is true only when the clearances of 
the two cylinders are proportionate. But the cylinders of stage 
compressors may, and usually do, have different clearances; so 
that the net high-pressure cylinder volume, at the end of re- 
expansion of the clearance air is not equal to that of the low-pres- 
sure or intake cylinder at the beginning of re-expansion. Never- 
theless, since the volume of air delivered by the low-pressure 
cylinder must necessarily be equal to the volume received by the 
high-pressure cylinder, disproportionate clearance does not affect 
the work done per unit of air compressed. The diagram of the 
high-pressure is merely displaced somewhat with respect to that 
of the low-pressure cylinder, as shown by Fig. 43, in which F E 
= F' E' and H D = H' D'. 

For disproportionate clearance, therefore, the work is expressed 
by the equation: 

^= ^1(V-^"')[(?) - -0- (V"-V)(f)*[ (.8) 

In equations (27) and (28) the symbols have the following sig- 
nifications : 

P and V = initial absolute pressure in pounds per square 

inch and initial volume in cubic feet. 
P'' and V = terminal absolute pressure and initial volume of 
air in the high-pressure cylinder. 
\'" = volume of there-expanded clearance air ( = J A 
in Fig. 42 and N A in Fig. 43). 

* For the sake of brevity, the steps in the deduction of the following work 
formulae are omitted. Those readers who desire to pursue the subject further will 
find a full discussion in Dr. Charles E. Lucke's forthcoming work on Engineering 
Thermodynamics. 



72 



COMPRESSED AIR PLANT 



The work and capacity of two-stage compressors with clear- 
ance may also be expressed in terms of displacement and volu- 
metric efficiency. 

Let Di and Dj = cylinder displacements, respectively of the 
low- and high-pressure cylinders. 




Fig. 43. — Reference Diagram. Work of Two-Stage Compressor, with Dispro- 
portionate Clearance. 



Ci and C2 = fractional clearances; whence C^ X D^ and 
C2 X D2 = the clearance volumes, and 
D, (i -I- CO and D2 (i -I- C2) = the total 
cyhnder volumes in cubic feet. 

A B 

E, = ^^ _ = volumetric efficiency of low-pressure cylinder. 



E,= 



NB 



y-= volumetric efficiency of high-pressure cylinder* 



Introducing these symbols: 

«.(m-^.-^.(?)"(i-T" -■](■■ w 



THE COMPRESSION OF AIR 



73 



which expresses in foot-pounds the work of the compressor in 
terms of the displacements and clearances of both cylinders. 
As the deduction is based on the assumption of best inter- 
cooler, or intermediate receiver, pressure, the displacements of 
the two cylinders must have a corresponding relation, to produce 
the most economical division of work betw^een them. This rela- 
tion in turn involves the clearances and the compression ratio. 

Referring again to Fig. 43, and denoting pressures and volumes 
by subscripts identical with the letters on the diagram: 

D, = H' D' -h H' O' = H D + H' O^ 



But H D = V^ - V, = V, -' - V, 



D, (I + Q) 



,p^^i -C.D, 



" "• (f) 

andH' O' = \v - V, = V,(|) »'- V,, = C, D,[(|^)r„-i] 



whence: D,=5iil±A) 



C, D, + C,D,[(^')/"- i] 



or. 



D,-C,D,[(|:)^-x]=D.[-i^'-C.] 

Vp / 









I + C, 



(3°) 



This equation expresses the ratio between the displacements 
of the low- and high-pressure cylinders, in terms of the initial and 
terminal pressures and the clearances. 

It may be desirable also to express the work in terms of the 
volumetric efficiencies. From the diagram: 

D,[n-C,-C,(f)^«] 



E,= 



Vt-Vc 

v.-v, 



Di 



[- c,-c,(£:)^.] 



74 COMPRESSED AIR PLANT 

and similarly: Eg = [_i + C2 - C2 (^j""" J 

Substituting E^ and E^ for their equivalents in equation (29) : 

The capacity, L, of the compressor is measured by the intake 
capacity of the low-pressure cylinder, or : 

L = V,-V„=D.[i + C, -C. (1^)^] .... (32) 

expressed in terms of displacement and clearance; and since 

Finally, if equation (29) be divided by equation (32) the work per 
unit of intake capacity of the compressor is expressed in terms of 
cylinder displacements and clearances. 

The subject of the performance of air compressors is presented 
in Chapter X. Tables are there given, showing the work actually 
required per foot of free air, for single-, two- and three-stage com- 
pression, including the work consumed by frictional and other 
losses. 



CHAPTER IV 

WET COMPRESSORS 

Although during the past fifteen years wet compressors have 
become almost obsolete in the United States, it is necessary to give 
some attention to them, not only because many are still used in 
P^urope, but also because a discussion of their design and operation 
will lead to a better understanding of the comparative merits of 
the systems of cooling employed in the modern dry compressors. 

Wet compressors ai'e of two kinds ; 

1. The so-called hydraulic-plunger compressors, in which water 
is introduced in bulk into the air cylinder, and is injected also in 
the form of spray, 

2. Those in which the cooling water is injected in the form 
of fine spray or jets only. 

Compressors of the first type comprise some of the earliest forms 
of air compressor. One of the best of this class is the modernized 
Dubois-Franfois, built at Seraing, Belgium. It has been widely 
used in Europe, for mining and tunnelling operations, and it is 
worth noting that, up to about 1877, one of them was employed at 
the Sutro tunnel, Nevada. Another well-known compressor of 
the same class, but of different design, is the Humboldt, made at 
Kalk, near Cologne, Germany. One of these also was erected 
at the Sutro tunnel, and did excellent work. A brief description 
of the old Humboldt compressor (Fig. 44) will serve to explain the 
principle and construction of these machines. 

The water constitutes a piston for compressing the air; an 
ordinary plunger, like that of a pump, moving in a horizontal 
cylinder filled with water. At each end of the cylinder, and con- 
nected with it by an easy curve, is a vertical air chamber. The 
upper ends of these chambers are provided with the necessary air 

75 



76 



COMPRESSED AIR PLANT 



inlet and discharge valves. As the piston reciprocates, the air 
is drawn alternately into one air chamber and compressed in the 
other, by the rise and fall of the water level. At the end of each 
stroke the air compressed by the rising mass of water in the air 
chamber passes through the discharge valves into the receiver. 
As the air is in contact with the water a partial cooling is effected, 
and to prevent the water itself from becoming heated a constant 
circulation must be maintained. A further cooling is brought 
about by the injection of sprays from a small force pump into the 
cylinder and vertical air chambers. The pump is operated from 




Fig. 44. — Humboldt Wet Compressor. 



the cross-head of the compressor itself. This type of compressor 
is simple, and if the sprays be copious the air is quite effectually 
cooled; but it is generally limited to rather slow speeds (only 100 
to 150 feet piston speed per minute or less in some cases), on ac- 
count of the inertia of the body of water. This is about one-third 
to two-fifths of the piston speed of modern dry compressors, and 
it follows that such engines are comparatively heavy and bulky for 
a given output of air, besides requiring expensive foundations. It 
is claimed, however, that a more recent form of Humboldt wet 
compressor can be run successfully at speeds of 300 to 360 feet per 
minute, the temperature of the air at discharge being kept at 77° to 
80° Fah.* This is such remarkably good work that the results are 
open to question, as far as regular, normal service is concerned. 
Lower speeds are certainly advisable for this form of compressor. 

* P. R. Bjorling, Colliery Guardian, Oct. 2d, 1896, pp. 629-630. 



WET COMPRESSORS 



77 



The machines are made of large size and are heavily and sub- 
stantially built. Violent shocks are apt to be caused by attempting 
to run at high speeds, for which reason the vertical air chambers 
join the cylinder with a curve of long radius to ease the movements 
of the mass of water. 

Fig. 45 shows a late type of the Hanarte wet compressor, many 
of which have been built for French and Belgian mines, and also 
for use in connection with ice-making plants. They are generally 
of large size, and are found to be highly efficient when run at piston 



■i""'i"'ii' III' '"■ 




Fig. 45. — Hanarte Wet Compressor. 

speeds of 250 to 275 feet per minute. The splayed out vertical 
ends of the cylinders cause the level of the water to rise slowly 
towards the end of the stroke, and afford space in the cylinder 
heads for large and readily accessible inlet and delivery valves. 
Sprays are used in addition to the water in bulk. 

A difficulty with wet compressors of this class is that an efficient 
circulation of cold water is not easy to maintain. Only a small 
quantity of fresh water can be injected at each stroke, and without 
copious sprays the cooling is imperfect. This is due to the fact 
that, although the mass of water kept in motion in the cylinder and 
air chambers is large, there is between it and the air only a surface 
contact. Since water is a poor conductor of heat, under these 
conditions it can hardly be questioned that the air is cooled more 
by contact with the relatively large area of the wet cylinder walls 
than bv its contact with the small superficial area of the rising and 



78 COMPRESSED AIR PLANT 

falling water. Another disadvantage is that the compressed air 
delivered from the cylinder is practically saturated with moisture. 

Compressors of the second class, in which the cooling water is 
used only in the form of jets or spray, constitute an improvement 
upon the older design, in being much less cumbrous and per- 
mitting a higher working speed. This method of cooling was first 
applied by CoUadon, at the St. Gothard tunnel. Though these 
compressors are still frequently used in Europe, they have given 
way in great measure to dry compressors, and in American practice 
have become almost obsolete. The air cylinder does not differ 
materially from that of the dry compressor. A small water pipe is 
tapped into each cylinder head and fine spray is injected in front 
of the piston while compression is taking place. 

Undoubtedly this system is superior to that involving the use of 
water in bulk. Since the water is in a state of fine division a rela- 
tively large surface of contact is presented, and the air is kept 
thoroughly saturated with moisture during compression. Zahner, 
in his " Transmission of Power by Compressed Air," p. 28, states 
that Colladon's St. Gothard compressors, "which were run at a 
piston speed of 345 feet, and compressed the air to an absolute 
tension of 8 atmospheres (103 pounds gauge pressure), gave an 
efficiency which never descended below 80 per cent, while the 
temperature of the air never rose higher than from 12° to 15° C. 
(53° to 59° F.)." The temperature of the injection water is not 
stated, but must have been very low to obtain such remarkable 
results. 

A dry compressor may be converted into a wet compressor 
merely by providing the cylinder with water jets. The injected 
water collects in the cylinder until enough is present to fill the piston 
clearance space at the end of the stroke. Then any additional 
amount of water is forced out at each stroke with the compressed 
air through the discharge valves into the air receiver. From the 
receiver the water is drained away from time to time. As the pis- 
ton clearance in well-designed compressors is extremely small, very 
little water can remain in the cylinder to be churned back and forth 
by the piston. The water used for injection should be as pure and 



WET COMPRESSORS 



79 



cold as possible. Gritty water must never be employed, as it 
would injure the cylinder, piston and valves. 

A proper injection apparatus should fulfil three conditions: 

1. The injection must commence at the beginning of the 
stroke and continue to the end, against the advancing piston. 

2. There should be a thorough diffusion of the water in the 
form of spray throughout the cylinder. By mere surface contact 
water takes up but little heat. Even a single strong jet is quite ef- 
fectual, however, because on striking the piston it is thoroughly 
broken into spray. 

3. A definite volume of water should be injected, the quantity in- 
creasing with the pressure under which the compressor is working; 
that is, with the quantity of heat generated. If the quantity of 
water used be insufficient to abstract the heat, a large amount of 
moisture is taken up by the warm air and carried into the receiver 
and piping. 

The heat units developed by compression having been cal- 
culated, the quantities of water required for different pressures 
are shovvn in the following table.* The average temperature of 
injection water may be taken as, say, 68° Fah., and is considered 

Table III 



P J^JIggTTW TTC 




Pounds of water to be in- 






Heat units 


jected at 68° F. to keep final 






developed by 

compression 

in one pound of 


temperature at 104° F. 




Gauge 




Above 






vacuum. 


pressure. 


free air. 


Per pound of 


Per cubic foot of 


Atmospheres. 


Pounds. 




free air. 


free air. 


2 


14.7 


58-310 


0-734 


0.056 


3 


29.4 


92.390 


1. 164 


0.089 


4 


44-1 


116.627 


1.469 


0.112 


5 


58.8 


135-388 


1. 701 


0.130 


6 


IZ-S 


151.700 


1. 891 


0.144 


7 


88.2 


163-735 


2.063 


0.158 


8 


102.9 


174-937 


2.204 


0.168 


9 


117. 6 


184.865 


2.329 


0.178 


10 


132-3 


193.701 


2.440 


0.186 


12 


161. 7 


209.090 


2.634 


0. 201 



* This table is taken in part from that given by Zahner, "Transmission of 
Power by Compressed Air," p. no, EngHsh units being substituted for French. 



8o 



COMPRESSED AIR PLANT 



as having accomplished its work if it leaves the cylinder at 104° 
Fah., these temperatures corresponding, respectively, to 20° and 
40° C. 

There is no practical advantage to be gained by using an 
excessive quantity of water, and care should be taken to inject no 
more than is required. The additional cooling effect of a greater 
mass of water in the cylinder would be but small — as has been re- 
marked under wet compressors of the first type — and more power 
would be consumed in pumping the water into the cylinder and 
then forcing it out again through the delivery valves. 



CHAPTER V 

DRY COMPRESSORS 

In the dry system of compression no water enters the air cylin- 
der except that which is carried as moisture in the air itself. All 
the cooling during compression, aside from radiation, is effected by 
a water envelope, or "jacket," surrounding the cylinder, and in 
which cold water is kept constantly circulating. 

Fig. 46 shows the longitudinal section of a Nordberg jacketed 
air cylinder. (Reference may also be made to Figs. 2, 5, 8, 11, 23, 
and other cuts of longitudinal sections, as illustrating different 
types of jacketed cylinders.) The cylinder is enclosed in an outer 
shell, leaving an annular space, J J, to be occupied by the water. 
Besides the annular jacket nearly one-half the area of each cylinder 
head is also covered by water jackets, K K. The remainder of the 
end areas is occupied by the suction and delivery valves, as shown. 
The air-delivery valves are sometimes placed radially, close to the 
cylinder ends, whereby a larger proportion of the area of the heads 
can be jacketed. This is true, for example, of one or two of the 
Laidlaw-Dunn-Gordon patterns. 

In Fig. 46 the circulation of water is elTected by pipes connecting 
with the openings A and B, respectively for inlet and discharge. 
To cause a proper circulation the spaces enclosed by the jacket are 
subdivided. The cold water enters at A, and after circulating 
through the annular and end jackets J J, K K, is finally discharged 
at B. The smaller jackets on the cylinder heads are designed to 
surround the valves and air passages as completely as possible, in 
order to exert the maximum degree of cooling. At C is a drain 
pipe through which the jacket is blown out occasionally to clear 
it of sediment. 

6 81 



d>2 



COMPRESSED AIR PLANT 



In some makes of compressor, the annular jacket is divided by 
vertical partitions, so that the cold water entering at the top 
passes first around about one-fifth of the length of the cylinder 
nearest each end. The water then circulates around the middle 
portion, and is discharged at the top. Although in this arrange^ 
ment the fact is recognized that at the end of the stroke, where the 
air pressure is highest, the greatest amount of heat is generated; 
still, in some of the same designs little, if any, of the cylinder-head 




f 

Fig. 46 — Air Cylinder of Nordberg Compressor. 

area is jacketed, because of the mode of placing the inlet and dis- 
charge valves. This would seem to be a defect because, on 
approaching the end of the stroke, the piston rapidly covers the 
annular jacket, leaving a very small part of its area available for 
cooling the hot compressed air while being discharged from the 
cylinder. It is at this point of the stroke that large end jackets ar 
most valuable. Similar provision for large jacket surfaces 
made in the Ingersoll-Rand compressors, particularly those of 
the ''Hurricane Inlet" type. The water passes first into the 
jackets on the cylinder heads and then successively through sev- 
eral separate compartments of the annular jacket. 



Dr 

le ■ 

i 



DRY COMPRESSORS 



83 




84 COMPRESSED AIR PLANT 

The jacket of one of the Laidlaw-Dunn-Gordon designs (Fig. 
47) is cast with eight longitudinal partitions, extending alternately 
from each end of the cylinder nearly to the opposite end. The 
water, which enters near the top, is forced to travel back and forth 
between the partitions and from one end of the cylinder to the other 
until it is finally discharged. An active circulation is thus main- 
tained. For furnishing the cooling water a tank is often provided, 
set at some elevation above the compressor, or a small pump may 
be employed. 

Naturally, a partial cooling only can be effected by water-jacket - 
ing the air cylinder. Much depends on the speed at which the com- 
pressor is run. In the best single-stage compression, to say seventy 
or seventy-five pounds, and at not over 300 feet piston speed, it is 
doubtful whether more than about one-half of the total possible 

. P' / V \ '^ 
cooling can be effected ; that is, in the equation — = ( — j ,n would 

be equal to, say, 1.22 to 1.25. Heat is generated faster than it can 
be abstracted, and only a portion of the volume of air passing 
through the cylinder comes into direct contact with the cooling sur- 
faces. It is important, therefore, that as much as possible of the 
total cylinder surface be covered by the jacket, and that the piston 
speed be moderate. But, in a dry compressor, as the air is com- 
paratively free from moisture, some heating is not so objectionable 
as it would be in a wet compressor. As a matter of fact, the cylin- 
der, discharge pipe, and even the receiver, are usually quite hot 
when the compressor is running at full speed ; often too hot to be 
touched with the hand. In a plant at Birmingham, England, with 
well-jacketed cylinders, and compressing only to forty-five pounds, 
a temperature of the air at delivery has been observed as high as 
280° F. In this case the compressor is large, so that the super- 
ficial area of the jackets is small as compared with the volume of 
the cylinder. It is probable that the heat of compression in 
dry compressors ranges from 200° to a maximum of 400° F. for 
the ordinary pressures used in mining, though it does not often 
exceed 350°. Care should be taken not to allow the temperature 



DRY COMPRESSORS 85 

to rise above this point.* At a large mine in Montana , the writer 
has observed the thin wrought-iron delivery pipe of a fifty-drill 
compressor red-hot for a distance of nearly six inches from the 
cylinder shell. Driving compressors at too high a speed (when not 
large enough for their work) is often the cause of the poor results 
complained of by some users of compressed air. 

In some compressors the inner shell of the air cylinder, i.e., 
between the cylinder and water-jacket, has been made of hard 
brass, which by its high conductivity assists in carrying off the 
heat. With the same end in view, the cylinder walls should be as 
thin as is consistent with safety. 

Besides its function of cooling the air during compression, the 
water-jacket of a dry compressor is indispensable from a mechanical 
point of view, in keeping down the temperature of the cylinder 
shell. Without some special provision for cooling the cylinder 
the metal vrould become hot enough to burn the oil, and render 
proper lubrication impossible. To furnish a larger cooling sur- 
face one of the older styles of Rand compressor had a hollow 
back piston-rod and hollow piston, through which water is circu- 
lated. To maintain circulation the back piston-rod worked 
telescopically in a stationary tube connected with the water 
supply. 

Piston Clearance in the Air Cylinder. In every engine, whether 
steam engine or compressor, the amount of clearance at the end of 
the stroke, between the piston and cylinder head, is a matter of 
some importance. It has a special bearing in the case of a dry 
compressor, which may be explained as follows. Toward the 
end of the stroke the compressed air in front of the piston begins 
to pass through the delivery valves as soon as its tension exceeds 
that of the air in the discharge pipe leading from the cylinder to 
the receiver. But remaining in the clearance space, on the com- 
pletion of the stroke, is a certain quantity of warm compressed 
air, which in the case of a dry compressor can never be discharged. 
On the back stroke the clearance air expands and partly fills the 
cylinder behind the piston. No air can enter through the inlet 
valves until the pressure inside the cylinder falls below atmos- 

*T. G. Lees, Trans. Federated Inst. Mining Engrs., Vol. XIV, p. 569. See 
also Chapter XIII of present volume. 



86 COMPRESSED AIR PLANT 

pheric pressure. It is never possible, therefore, to take a full 
cylinder of fresh air even under the best conditions, and the clear- 
ance space must be made as small as possible, say, about one- 
sixteenth inch. Or, the clearance may be expressed as a ratio, by 
dividing the clearance volume by the entire cylinder volume swept 
through by the piston in making its stroke. In a wet compressor 
the clearance space is filled with water, and therefore does not 
produce the effect just described. 

In cylinders of the same diameter and having the same amount 
of linear clearance at the ends of the stroke, it is evident that the 
ratio between cylinder volume and clearance volume depends on 
the length of stroke. This ratio is generally largest in short- 
stroke compressors and smallest in those of long stroke. It 
varies, also, in compressors of different makers. As examples 
may be cited the following ratios between cylinder and clear- 
ance volumes of several Ingersoll-Rand compressors, of different 
strokes: 

14 inch stroke 0190 

21 " " 0176 

24 '' " 0126 

36 " " 0112 

48 " " .0093 

ranging thus from about 2 per cent, down to i per cent. Some 
recent compressors, built by the same company, of 4 2 -inch 
stroke, but of relatively small cylinder diameters, have piston 
clearances as small as .78, .80 and .90 of one per cent., and 
several of 36-inch stroke have clearances of .83 and .84 of 
one per cent. In a recent type of the Leyner compressor, 
with a 22-inch cylinder, this percentage is stated to be 1.02, 
and in several of the Laidlaw-Dunn-Gordon compressors of 
standard type, it ranges from .75 to 1.25 per cent. Clearances 
are generally larger, however; thus, in a new design of direct, 
electric-driven, two-stage compressors, of the Ingersoll-Rand 
Co., the following clearances are found in the low-pressure 
cvlinders : 



DRY COMPRESSORS 



87 



28 inch X 24 inch 2.15 per cent. 

23 " X 20 inch 1.70 " "- 

19 " X 16 '^ 1.85 '' '' 

18 " X 14 " 2.00 '' 

17 '' X 14 " 2.19 " " 

The compressors of some other makes have clearances as high as 
2 J per cent, of the cylinder volume. 

With few exceptions the lowest figures apply to large, long- 




Air Cai-d showing 
effect oLclearance. 
Voluma between 6 and c 



■[-Atmospheric 
5 [C Line 

f i — Vacuum 

Fig. 48. 



stroke compressors; the higher to the small, short-stroke ma- 
chines in common use for many kinds of service. 

The diagram, Fig. 48, shows the effect of clearance. Before the 
inlet valves can open, the piston must travel from c to b, and the 
corresponding cylinder volume passed through by the piston repre- 
sents the percentage of loss of volumetric capacity as stated above. 
The actual effect of piston clearance on the volumetric efficiency of 
the compressor of course depends on the number of compressions ; 
that is, on the air pressure produced. The higher the terminal 
pressure, the farther must the piston travel before the inlet valves 
can open and the greater is the distance from c to b, in the diagram. 
It may be added, however, that this reduction of capacity, although 
a matter of considerable importance in the operation of the com- 
pressor, does not involve a corresponding loss of useful work. 
The compressed air remaining in the clearance space helps to 



88 



COMPRESSED AIR PLANT 



overcome the inertia of the moving parts at the beginning of the 
return stroke, and to compress the air on the other side of the 



UbO 
825 


1 






1 




/ 


/ 




/ 


/ 


/ 




/ 




1 




/ 


/ 






/ 




175 
150 
125 
100 
75 
50 
25 









/ 


/ 


/ 


/ 


/ 










/ , 


5/ 


/ 


/ 


/ 




/ 








J 


/ 


y 


/ 


/ y 












/ 


/ 


Ay ^ 




Y 












/ 


// 






















/// 


Y/ 


















/m 


f 




















P^ 























5 10 15 20 25 30 35 40 45 50 

Per Cent of Piston Displacement 

Fig. 49- 

piston. A part of the work expended in comipressing the clearance 
air is thus recovered. It has been observed that the clearance air 
cools slightly during the momentary stoppage of the piston as the 
stroke is reversed, but the consequent reduction of pressure is a 
negligible quantity. In expanding behind the retreating piston^ 
however, the clearance air rapidly gives up its heat and does not, 
therefore, tend to raise the temperature of the incoming atmos- 
pheric air. 

The effect of piston clearance in reducing the capacity of a dry 
compressor is shown clearly by the diagram, Fig. 49, which is 



DRY COMPRESSORS 



89 



reproduced here by kind permission from Engineering News, May 
30th, 1 901. It shows that, for clearances above one per cent, the 
loss becomes serious even at pressures of seventy-five to one hun- 
dred pounds. 

Fig. 50 indicates the method of reducing the clearance for or- 
dinar}^ pistons, by casting a recess in the cylinder head to receive 
the projecting piston nut at the end of the stroke. The loss of 
volumetric capacity due to clearance of course increases with 
the air pressure, and in some compressors the piston is run exceed- 



1^^^^^^^ 





Fig. 50. 



Fig. 51. 



ingly close to the cylinder head. When this is the case the com- 
pressor must have careful attention, so that if the working length of 
the connecting rod should be varied in fitting new brasses, the pis- 
ton will not be in danger of striking the cylinder head. 

The Johnson compressor, made in England, has an ingeniously 
designed piston (Fig. 51) to meet the difficulty just mentioned. It 
is composed of two disks, c and d, mounted on a brass sleeve, 
screwed on the piston-rod, h, and held in place by collar and lock- 
nut. The disks are so cast as to leave between them a recess, in 
which is placed a heavy helical spring, /. This spring is compressed 
sufficiently between the disks to prevent it from being further 
compressed under the maximum working air pressure, but the 



90 COMPRESSED AIR PLANT 

clearance at the ends of the stroke is extremely small, and should 
the piston strike the cylinder head the spring gives slightly and an 
injurious shock is avoided.* 

A number of other devices have been adopted for overcoming 
the disadvantages of piston clearance. Two examples may bfe 
given : 

1. Longitudinal bye-pass grooves (B B) are cast in the inner 
surface of the cylinder near the ends, Fig. 50, so that when the 
piston reaches the end of its stroke a part of these grooves is un- 
covered, and the compressed air in the clearance space passes to 
the other side of the piston. 

2. In slide-valve compressors the valve may be provided with 
a so-called "trick-passage." At the end of the stroke this passage 
is brought into connection with two small ports entering the ex- 
treme ends of the cylinder. Through these passages the high- 
pressure air in the clearance space is released into the other end 
of the cylinder. 

Although by these methods the released air becomes of direct 
benefit, there is a decided objection to their employment if all the 
confined air be allowed to pass over, because the heavy pressure 
on the piston is suddenly removed, and there is a shock to the mov- 
ing parts which is clearly evidenced by pounding at the end of the 
stroke. In the most recent forms of compressor made in the 
United States the clearance space is very small, but the air con- 
fined in it is not released. 

Dry Versus Wet Compression 

Up to about 1885 there seemed to be little doubt among me- 
chanical engineers that the wet compressors are, upon the whole, 
superior to the dry, because by bringing the air into direct contact 
with water the heat is most effectually absorbed. This view is 
correct so far as heat loss alone is concerned, provided the water 
introduced into the cylinder is properly applied, as pointed out in 

* Bjorling, Colliery Guardian, Aug. 7th, 1896, p. 272. 



DRY COMPRESSORS 9 1 

Chapter IV. Without cooling the percentage of work converted 
into heat during compression, and therefore lost, is as follows : 



Comp 



ession to 2 atmospheres, 9.2 % loss. 



15-0 % 
19-6 % 

21.3 % 
24.0 % 
26.0 % 
27-4 % 



In well-designed dry compressors, working at a pressure of 5 
atmospheres, the heat loss is reduced about one-half, or from 
21.3 per cent, to 11 per cent. Frequently, however, in ordinary 
mining practice, with single-stage compressors, the loss is fully 
15 per cent. By spray injection this loss has been cut down in the 
best American practice to as little as 3.6 per cent.,* and in some 
of the large, slow-running European wet compressors to 1.6 per 
cent. But the question of heat loss is not the only consideration. 
Low first cost and simplicity of construction are often more ad- 
vantageous than a close approximation to isothermal compression. 
Latterly the wet systems have lost ground, and it is probable that 
no wet compressors are now being built in the United States. In 
Europe also dry compression has grown in favor, at least ft)r min- 
ing plants and others of moderate size. The matter may be con- 
sidered from two standpoints, as regards: 

1. The effect of injected water upon the compressed air and 
the machines using it. 

2. The effect of the water upon the working of the compressor. 
In addition, it is necessary to take account of the relative efficien- 
cies of the two types, but this will be deferred until later. 

Firstj it is unquestionable that by using large slow-speed en- 
gines, and an abundance of injection water, the air is well cooled, 
though at a higher first cost for plant. Wet compression gives a 
good indicator card. It is shown by Table IV that in compressing 
moist air somewhat less work is expended than for dry air. This 

* As stated regarding the old Ingersoll injection compressor, by W. L. Saunders, 
" Compressed Air Production," p. 24. 



92 



COMPRESSED AIR PLANT 



is due to the fact that the specific heat of watery vapor is about 
twice that of dry air; therefore in the presence of moisture more 
heat is required to raise the temperature of the air in the com- 
pressing cylinder, and the loss of work from this cause is reduced. 

Table IV 



Absolute 
Pressure. 


Gauge 
Pressure. 
Pounds. 


Foot Pounds of Work Required to Compress 
One Pound of Air. 


Atmospheres. 


Dry Compression. 


With Sufifiicient 
Moisture. 


I 

2 

3 
4 
5 
6 


O 

14-7 
29.4 

44-1 
58.8 


23,500 
37,000 
48,500 
58,500 
67,000 
75,000 


22,500 
35,000 
45,000 
52,500 
60,000 
66,000 



Theoretically, a corresponding economy takes place also when 
the air is expanded again in the machine using it. 

Notwithstanding these advantages, several serious objections 
became apparent in the use of the wet system of compression. 
Other things being equal, the amount of heat given up during 
compression is proportional to the difference of temperature be- 
tween the air taken into the cylinder and the injected water, and to 
the time of contact between the air and water. Under ordinary 
circumstances this difference of temperature is zero at the beginning 
of the stroke, reaching its maximum at the end. It follows: (i) 
that to attain a fair approach to isothermal compression the piston 
speed must be very slow; (2) that during the first part of the stroke 
but little heat is removed, and it is only when compression is com- 
plete, and the air begins to pass from the cylinder through the 
discharge valves, that the cooling effect is at its maximum. At 
ordinary piston speeds, therefore, a large proportion of the total 
heat must be given up after the discharge valves have opened ; in 
other words, after compression is completed. For this reason it 
would appear that, so far as economy of work is concerned, the 
lower final temperature due to spray injection is in a measure de- 



DRY COMPRESSORS 93 

ceptive. The warmth of the air at discharge augments its moisture- 
carrying capacity, and though it is intended that the separation of 
the water shall be as complete as possible in the air receiver, still 
it must of necessity be imperfect in a receiver of any reasonable 
size. Much moisture passes into the air mains, and deposits as 
the air cools down in long lines of piping. In cold weather it may 
freeze so as to reduce the effective diameter of the pipe. The mois- 
ture remaining in the air has a further ill effect when it is used. 
At the instant of exhaust by the drill, or other air engine, the in- 
tense cold produced by expansion causes the formation of trouble- 
some accumulations of ice in the exhaust passages. 

As to the dry compressor it must be admitted that as air is a 
poor conductor of heat it has little opportunity to give up its heat 
of compression between the strokes of the piston. Besides this, 
the piston, as it advances, rapidly covers the jacket-cooled sur- 
face of the cylinder. However, although atmospheric air as taken 
into the compressor always contains moisture, which will make its 
appearance as frost at the exhaust of the air machine, still there is 
not enough of it to cause serious trouble.* The deliver}- of warm 
air by a dry compressor is far less objectionable than warm air 
from a wet compressor. 

Second, as to the efTect of injected water upon the working of 
the compressor. Under the best of circumstances water in the air 
cylinder is objectionable, because it makes lubrication difhcult, 
causes rust, and increasing the wear of piston and cylinder in- 
volves greater expense for repairs and renewal of parts. No sat- 
isfactory method has ever been devised for lubricating the inner 
surface of wet compressor cylinders. This is one of the chief diffi- 
culties with wet compressors, and becomes most serious when the 
water is impure or gritty. It must, of course, contain no trace of 
acid, such as is often present in mine water. Water that is com- 

* The quantity of moisture in the atmosphere, or its humidity, varies with the 
climate, the season of the year, and in a measure with the altitude above sea-level. 
It is usually greatest near the ocean or any large body of water. What is commonly 
called dry atmospheric air contains from forty to fifty per cent, of the quantity neces- 
sary to saturate it. The degree of saturation in summer often reaches ninety per 
cent, or more. 



94 COMPRESSED AIR PLANT 

paratively harmless for use in jackets might be decidedly injuri- 
ous to the finished surfaces of working parts. It has been stated 
by Mr. W. L. Saunders that, although the thermal loss is higher 
in dry than wet compressors, the frictional loss in the moving 
parts is considerably higher in the wet compressor. The net 
economy of the best wet compressors is probably no greater than 
that of the best American dry compressors. 

It is urged on behalf of wet compression that the piston-clear- 
ance space is filled with water, and the capacity of the compressor 
is therefore increased. While this is true, yet, as water is in- 
compressible, and as a part of it must be forced out through 
the discharge valves at each stroke, the wet compressor is com- 
pelled to work in a measure like a water pump. Furthermore, 
closer attendance is required to regulate the water supply. The 
drip cock at the bottom of the receiver must also be watched more 
closely to prevent flooding, and there is the disadvantage of having 
an injection pump to care for and regulate. 



CHAPTER VI 

COMPOUND OR STAGE COMPRESSORS 

Compound or stage compressors have two or more air cyl- 
inders, between which the total work of compression is divided. 
The air cylinders are placed tandem on a common piston-rod, as in 
straight-line machines, or respectively tandem with the steam cyl- 
inders in the duplex type. In two-stage compressors air at at- 
mospheric pressure is taken into the large or low-pressure cylinder; 
is there compressed to a certain point, and is then forced into the 
smaller or high-pressure cylinder, where it is brought up to the 
required tension (see Figs. 8 to 23). Manifestly, the size of the 
low-pressure or intake cylinder determines the capacity of the com- 
pressor. In a certain sense, the operation of a two-stage com- 
I pressor is the reverse of that of a compound steam engine. 

The theory and application of stage compression are readily 
comprehended. Since the heat of compression increases with the 
air pressure produced — though not proportionately, as has been 
shown — it follows that the higher the pressure the more difficult 
does it become to keep down the temperature to a point permutting 
efficient operation of the compressor and proper lubrication of the 
air cylinder. In attempting, with a single-cylinder dry compressor, 
to compress even to 90 pounds gauge, the theoretical final cylinder 
temperature becomes 459° F., and at 100 pounds gauge 485° F. 
Though some heat is dissipated by radiation, the actual working 
temperatures corresponding to these pressures are still too high 
to be dealt with effectually by the ordinary water-jacket, because in 
a single cylinder the superficial area to which cooling can be applied 
is too small relatively to the volume of air, and the total compression 
period too short. Even when working at moderate piston speeds, 
say, not over 350 to 400 feet per minute, the cooling is very 

95 



96 COMPRESSED AIR PLANT 

imperfect. The compressed air, as discharged from the cylinder, 
is still hot, so that considerable loss of pressure and of work, due 
to subsequent cooling, is inevitable. 

These disadvantages are in large measure overcome by the 
adoption of stage compression, and, in view of the fact that this 
system was introduced over twenty-five years ago, it would appear 
strange that until quite recently it has been neglected, by nearly 
all compressor builders, for the ordinary pressures used in mining, 
tunnelling, and similar work. 

Formerly it was customary to employ stage compression 
only when high pressures were required, such as for pneumatic 
locomotives, riveting machines, presses, compression of gases, 
pneumatic guns, etc. For such service stage compression is in- 
dispensable; and the higher the pressure the greater becomes the 
necessity for compounding the air cylinders and the comparative 
efficiency of the system. To produce very high pressures, of 500 
to 1,000 pounds or more, three- and four-stage compression is 
employed. 

But it is now generally recognized that two-stage compression 
when properly applied presents some advantages even for press- 
ures of seventy to eighty pounds, as commonly adopted for ma- 
chine drills and ordinary air engines. The cooling during com- 
pression is more thorough because the total heat generated is 
divided between two or more cylinders. In each cylinder the tem- 
perature is lower than when the same total pressure is produced in 
a single cylinder, and the combined water-jackets afford a much 
larger cooling surface. 

A further cooling is effected by an "intercooler," placed be- 
tween the cylinders. This constitutes one of the most important fea- 
tures of stage compression. It is an intermediate cooling-chamber, 
through which the partially compressed air from the intake or low- 
pressure cylinder passes on its way to the high-pressure cylinder. 
The temperature of the air is here reduced, so that when the 
high-pressure piston begins its work the temperature of the volume 
of air on which it acts is considerably below that at which the air 
was discharged from the low-pressure cylinder. Obviously, the 



COMPOUND OR STAGE COMPRESSORS 97 

total reduction of temperature effected depends on the volume 
of the air under compression, the area of the cooling surfaces and 
the length of time the air is in contact with these surfaces; or, in 
other words, on the piston speed. The construction of the inter- 
cooler will be taken up later. 

It should not be inferred from what precedes that stage com- 
pression per se is always applicable, nor that it is necessarily more 
economical than compression in a single cylinder. Concerning 
this, several fairly well defined, though interrelated statements 
may here be made: 

1. Although stage compression is theoretically advantageous for 
all pressures, it becomes of doubtful utility for gauge pressures of 
much less than seventy-five pounds, because of the small saving 
as compared with the greater first cost and running expenses of 
the more complicated mechanism. It is generally applicable for 
pressures higher than seventy to seventy-five pounds. 

2. Stage compression is specially useful for large compressors, 
in which the percentage of saving will represent an amount suffi- 
cient to vv^arrant the greater first cost of plant. 

3. The higher thermodynamic efficiency of stage compression is 
in some degree offset, and in poorly designed plants may be entirely 
neutrali;ied, by the increased frictional losses involved in the use of 
several cylinders. In other words, when employing stage com- 
pression, advantage should always be taken of the opportunity 
to use a well-designed, economically working steam end, together 
with large and efficient cooling arrangements for the air end. 
If these requirements be not fulfilled, stage compression may easily 
cost more per cubic foot of air delivered than simple compression 
by a properly designed compressor. 

Almost all stage compressors are double-acting; that is, on each 
forward and back stroke air is taken into the cylinders on one side 
of the piston, while compression and deliver}' are going on on the 
other side. The operation of the single-acting form, occasionally 
employed, will be considered first. It is materially different from 
that of the double acting compressor, but its description will aid 
in setting forth the subject of stage compression. 
7 



98 COMPRESSED AIR PLANT 

Single-Acting Two-Stage Compressor. Supposing the intake, or 
low-pressure, cylinder to be filled with free air just taken in, the 
advancing piston compresses the air until a point somewhat beyond 
half stroke is reached. At this point the delivery valves open, and 
during the remainder of the stroke the compressed air, at, say, 
thirty to thirty-five pounds pressure, is being forced out through 
the connecting pipe and passages into the second or high-pressure 
cylinder. Meanwhile, no work is being done by the high-pressure 
piston. On the return stroke the air at the low pressure which 
was delivered into the high -pressure cylinder is compressed to 
the required final tension and discharged. During this return 
stroke no work is done in the low-pressure cylinder, except that 
another charge of free air is drawn in. Thus, the intake stroke 
of the low-pressure cylinder is the compression and delivery stroke 
of the high-pressure, and vice versa. During the low-pressure 
intake stroke the portion of partly compressed air remaining in 
the pipe or passage connecting the cylinders is unaffected, as it is 
shut off from both cylinders by the valves at either end. At the 
beginning of the return stroke of the high-pressure cylinder the 
air in the connecting pipe begins to flow into this cylinder, and its 
pressure diminishes according to the relative volumes of pipe and 
cylinder. In the mean time the air is being compressed in the 
low-pressure cylinder, and when its tension exceeds that in the 
connecting pipe (that is, at, say, half stroke) it begins to pass 
through the delivery valves into this pipe. During the remainder 
of the stroke the low-pressure piston is in reality acting upon 
and compressing, not only the air in its own cylinder, but also 
that which is m the connecting pipe and high-pressure cylinder. 

A serious disadvantage of the single-acting, two-stage compress- 
or of this form is that the net resistances in the two cylinders are not 
equalized. Although the actual work of compression is designed 
to be the same in both cylinders, equalization of the resistances 
throughout both strokes is practically impossible because, in the 
second half of the forward stroke of the intake piston, the air de- 
livered by it acts as a back pressure on the high-pressure piston, 
which is travelling in the same direction. This back pressure,^ in 



COMPOUND OR STAGE COMPRESSORS 99 

turn, assists the movement of the loAv-pressure piston during its 
compression stroke. In this stroke, therefore, less total resistance 
is presented than during the compression stroke of the high -press- 
ure piston. It has been pointed out by Mr. Frank Richards that, 
''to decrease the diameter of the high-pressure cylinder would 
tend toward equalization of the resistances, by allowing the intake 
cylinder to do more work, and compress the air to a higher press- 
ure; but to raise the pressure (at delivery) in this cylinder would 
be to defeat the object of two-stage compression — that of allowing 
an efficient cooling of the air, and a reduction of its volume before 
its compression is too far advanced." In stage compression it is 
a fundamental principle that the cylinders should be so propor- 
tioned that the total work is divided equally between them. This 
secures the largest saving possible in the mechanical work of the 
compressor, as well as in efficiency of the cooling apparatus. 

Double- Acting Two-Stage Compressors. The operation of this 
type is more satisfactory than that of the single-acting two-stage 
compressor, because, firstj the cycle of operations during each 
forward and back stroke is the same; d^nd, second, iht distribu- 
tion of the resistances throughout the stroke may be made more 
uniform. 

A number of combinations in the arrangement of the steam and 
air cylinders are possible, but only three forms need to be noticed, 
as representing accepted practice, viz.: the straight-line, two-stage 
compressor (Figs. 7 to 12) and the duplex forms, consisting of a 
pair of cross-compound air cylinders, placed tandem to either 
twin-simple, or cross-compound steam cylinders (Fig. 14 to 23). 
The last-named is undoubtedly the best for large plants. 

The principles of the mode of operation of all three designs may 
be illustrated by reference to Fig. 52, which shows diagrammat- 
ically a Nor^valk two-stage straight-line compressor. 

Assuming that the pistons have reached the end of their forward 
stroke, the conditions in the two cylinders are approximately as 
follows: The low-pressure cylinder (D) is full of air, practically 
at atmospheric pressure, while the high-pressure cylinder (G), to- 
gether with the intercooler (F) and connecting passages, are oc- 



lOO 



COMPRESSED AIR PLANT 




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COMPOUND OR STAGE COMPRESSORS lOI 

cupied by air just delivered from the low-pressure cylinder, at, say, 
30 to 35 pounds, or something less than one-half the final pressure. 
On the reverse stroke the free air in front of the low-pressure piston 
is compressed to 30 pounds and delivered into the intercoolcr 
and high-pressure cylinder, while the air already occupying the lat- 
ter is brought up to the final pressure and discharged. This must 
be considered only as a rough description of what takes place in 
the air cylinders during a complete forward and back stroke. 

As usually constructed for standard tandem, two-stage com- 
pressors, the volumetric capacities of the low- and high-pressure 
air cylinders are to each other in the ratio of about ten to four. 
The intention is to proportion the two cylinders so that their 
ratios of compression are nearly equal. Thus the distribution 
of work and the heat generated in the cylinders will be equalized 
and most effectually dealt with by the intercooler, provided 
the latter properly performs its functions. Practice as regards 
the relative volume of the intercooler and cylinders has not 
yet been completely standardized. It has undergone consid- 
erable change in the past few years. As clearer conceptions 
have been reached of the fundamentally important functions of the 
intercooler in stage compression, and in recognition of the fact 
that the first cost of even a very large intercooler is moderate, while 
its running expenses are practically nil, the tendency now is to make 
it of much greater volumetric capacity than formerly. Such in- 
crease of size produces substantial gain in thermodynamic efficiency. 
The hot compressed air delivered by the low-pressure cylinder is 
kept longer in contact with the cooling surfaces because of its 
reduced speed of flow through the larger cross-sectional area of the 
intercooler, and it enters the high-pressure cylinder, to undergo the 
second stage of compression, with a temperature that may readily 
be made to approximate closely to the normal. On the other hand, 
it is clear that the connecting passages between the cylinders and 
intercooler should be of as small volume as is consistent with free- 
dom from excessive frictional resistance in the flow of the air 
through them ; because the air occupying these passages at any given 
time is exposed to but little cooling save that due to radiation. 



I02 COMPRESSED AIR PLANT 

With these points in view, it may be assumed in good practice 
that, if the volume of the low-pressure cylinder be taken as lO, then 
the volume of its connection with the intercooler should be, say, 1.5, 
of the intercooler 4, of the connection to the high-pressure cylin- 
der 1.5, and of the high-pressure cylinder 4. (It may be noted 
that there is no reason why the net capacity of the intercooler 
should not be even greater than is here assumed.) Having these 
proportionate volumetric capacities, the following sequence of 
operations will take place while the compressor is making a single 
stroke. Suppose this stroke to be from right to left, as indicated 
by the arrows in Fig. 52. 

By the previous stroke (from left to right) the intercooler and 
both of its connections to the cylinders, representing a volume = 1.5 
4- 4 + 1.5, were filled with air compressed, at, say thirty pounds. 
This body of air was then shut off from both cylinders by their 
respective valves, and has lost part of its heat and pressure by the 
action of the intercooler. After reversal, and during the first part 
of the following (left-hand) stroke, the low-pressure piston acts only 
on the cylinderful of free air just taken in (volume =10).* While 
this is being compressed, the advance of the high-pressure piston 
causes the compressed air already in the intercooler and its con- 
nections to begin to flow into the high-pressure cylinder, thereby 
increasing in volume and decreasing in pressure, until a point, say, 
a little beyond mid-stroke is reached. On passing this point the 
air pressure in front of the low-pressure piston rises slightly higher 
than that in the intercooler and the corresponding low-pressure 
delivery valves open, so that the low-pressure piston acts upon the 

entire body of air — volume = hi.5 + 4+i-5+— = 14. Then, 

2 2 

until the end of the stroke, both cylinders are in communication 
through the intercooler, i.e., from the left-hand end of the low- 
pressure cylinder to the right-hand end of the high-pressure cylin- 

* The general method of analysis here given is similar to that employed some 
years ago by Frank Richards, " Compressed Air," pp. 86-87, though the quantities 
used are taken to represent a closer approach to current practice in the proportions 
of the parts. 



I 



COMPOUND OR STAGE COMPRESSORS IO3 

der, as shown by the arrows in the cut, and an approximate equal- 
ization of pressure is established throughout. 

Up to the time of the opening of the left-hand, low-pressure 
delivery valves, the air in the intercooler, and still under its in- 
fluence, has been isolated from the low-pressure cylinder, in which 
compression has progressed without other cooling than that effected 
by the cylinder water-jacket. But when the warm, partly com- 
pressed air begins to pass from the low-pressure into the high- 
pressure cylinder, through the intercooler, the influence of the 
latter is exerted upon a new body of air. At the end of the left- 
hand stroke the closing of the delivery valves again shuts off the 
air in the intercooler from both cylinders. The high-pressure cylin- 
der, on the right-hand side of the piston, is occupied by a body of 
air whose temperature has been reduced by the combined effect of 
the intercooler and both water-jackets to a point much below that 
due to the working pressure of the low-pressure cylinder, and 
whose pressure has dropped correspondingly. 

Now, in the latter part of the left-hand stroke, when the low 
pressure delivery valves have opened and the piston of this cylin- 
der is acting on the volume 14, as stated above, a portion of this 

2 -|-i.5 
air (volume = =25 percent.) of the total has passed beyond 

the influence of the intercooler, and another portion (volume = 

— =46 per cent.) has not yet reached it. A similar statement 

of the distribution of the air with respect to the intercooler may be 
made for other points of the stroke. At the end of the left-hand 
stroke under consideration the volume of compressed air in the 
low-pressure cylinder = 0, in the intercooler and its connections 
1.5 -f 4 + 1.5 = 7, ^^d i^ the high-pressure cylinder 4, a total of 
II, of which 1.5 has not reached the intercooler but has been 
affected only by the low-pressure water-jacket. 

This analysis should be clearly understood in forming a correct 
estimate of the work actually accomplished by the intercooler. 
It emphasizes the importance not only of employing an inter- 
cooler whose volumetric capacity is large relatively to the cylinders, 



I04 COMPRESSED AIR PLANT 

but also of making the conDecting passages small. It is evident 
that one-half of the total \\x)rk of compression — that performed in 
the high-pressure cylinder^ — is done solely under such cooling in- 
fluence as may be exerted by the water-jackets of this cylinder. 
The jackets of both cylinders should, therefore, be as large in area 
as possible, with an efficient circulation of cold water. They 
should cover not merely the cylinder barrels, but as much of the 
heads as the spaces occupied by the valves will permit. In the 
latter respect some recent compressor designs are deficient. 

The details of the distribution of the air in the foregoing de- 
scription apply exactly only to compressors in which the air cyl- 
inders are tandem to each other. In the duplex stage-compressors, 
where the air cylinders usually are, and always should be, cross- 
compounded, the cycle of operations is different because the pis- 
tons, instead of m^oving together in the same direction, work with 
one crank 90° in advance of the other. 

As stated above, it is intended in stage compression that the 
total work done shall be equally divided between the air cylinders. 
But, by reason of the frequent variations in receiver pressure, up- 
on which depends the actual terminal pressure of the high-press- 
ure cylinder, an approximate equalization only can be attained 
in practice. On the basis of some terminal pressure taken as 
normal, such diameters are assigned to the cylinders as will make 
their compression ratios equal, or nearly so. Take, for example, 
a pair of cylinders, 15 ins. and 24 ins. in diameter, to produce 
a final pressure of 85 lbs. gauge. x\ssuming that the air between 
the stages is cooled to the original temperature, the absolute in- 
take pressures of the cylinders will be inversely proportional 
to the squares of their diameters, or: 15^ : 24^ : : 14.7 : 37.64, 
The absolute pressure of 37.64 lbs., as delivered by the low- 
pressure cylinder, is theoretically equal to the intake pressure of 
the high-pressure cylinder. The ratio of compression in the low- 
pressure cylinder is: — ]r~ =0-3905; and in the high-pressure 
cylinder: — — =0.3775. This would be quite as close to per- 
fect equalization as is necessary. 



COMPOUND OR STAGE COMPRESSORS 



lo; 



Construction of the Intercooler. A number of forms are novv^ 
in use. As commonly constructed for straight-line compressors, 
the intercooler consists of a long cylindrical chamber, containing 
a number of parallel, thin brass (sometimes wrought-iron) tubes, 
through which cold water is circulated. The air to be cooled 
passes through the spaces between the tubes. The intercooler 
is placed in a convenient position between and above tlie cylinders, 
and as close to them as possible, so that the connecting passages 
may be short and of small volume. As already stated, the air con- 
tained in these passages at any given time is denied the cooling 




Fig. 53. — Horizontal Intercooler. Ingersoll-Rand Co. 

effect both of the cylinder water-jackets and of the intercooler 
itself. In Fig. 52 the intercooler is indicated at F: in Fig. 8 the 
Norwalk intercooler is shown in longitudinal section. Fig. 53 
illustrates a large horizontal intercooler, as built by the Ingersoll- 
Rand Co. x^nother design, for cross-compound air cylinders, by 
the Sullivan Machinery Co., is shown in Fig. 54, a large intercooler 
being placed crosswise below the cylinders. In many cross-com- 
pound compressors the intercooler is mounted above the cylinders. 
The tendency now is to increase the size and volume of the inter- 
cooling chamber, relatively to the volume of the cylinders. 

The air delivered from the low-pressure cylinder passes on its 
way to the high-pressure cylinder betw^een the intercooler tubes, 
which must be sufficiently close together thoroughly to split up the 
body of air traversing the intermediate spaces and so secure the 
maximum cooling effect. It is intended that the temperature of 



io6 



COMPRESSED AIR PLANT 




COMPOUND OR STAGE COMPRESSORS 107 

the air, on leaving the intercooler and entering the high-pressure 
cyhnder, shall be reduced nearly to the normal. The effect of this 
drop in temperature upon the compression curve of a two-stage 
compressor is shown by Fig. 57; the cur^•e of the high-press- 
ure cylinder should, and often does, begin close to the iso- 
thermal line. 

In the construction of the intercooler brass tubes are perhaps 
preferable to those of iron because of their higher conductivity; 
but, on the other hand, iron tubes cost less, and on account of their 
greater roughness present a larger cooling surface to the air flow- 
ing between them. In either case they should be as thin as is con- 
sistent with the necessary strength. The tubes are expanded into 
tube-sheets at each end, and by means of two or more baffle- 
plates, set equidistant between the ends, the air is compelled to 
pass through the entire volume of the intercooler. The water- 
heads at the ends are so divided that the water is caused to circulate 
actively back and forth several times, before it is finally discharged, 
as shown by the small arrows in Fig. 53. For convenience the 
water supply is usually connected with the circulating system of the 
cylinder-jackets. 

Fig. 55 illustrates a peculiar system of intercooling adopted in 
the Leyner compressor. A number of horizontal iron or bronze 
tubes are enclosed in the annular water-jacket spaces, between the 
inner and outer shells of the cylinder. The piston being at the 
middle point of its stroke, the inlet valves at the left-hand end of 
the low-pressure cylinder are open and taking in air. Meantime 
the air in front of the piston, having been compressed, is passing 
out through the deliver}^ valves into the air chamber or head at 
the right-hand end of the cylinder. This air is thence forced by 
horizontal baffle-plates in the air chamber through the upper set 
of intercooler tubes, and into the left-hand end of the cylinder. 
It flows next to the right, through the lower set of intercooler tubes, 
and as sho^\Ti by the arrows enters the lower tubes of the high-press- 
ure cylinder. From the right-hand air head of this cylinder the 
air is directed by baffle-plates back through the upper set of tubes to 
the left-hand end of the high-pressure cylinder, into which it enters 



I08 COMPRESSED AIR PLANT 

through the corresponding inlet valves. The air already com- 
pressed in this cylinder is shown as passing through the large upper 
aftercooling tubes to its own air chamber, which leads to the dis- 
charge pipe. It will be noted that the low-pressure air, in being 
subdivided into small volumes and compelled to change its direc- 
tion several times in passing back and forth through the intercooler 
tubes, is well cooled before entering the high- pressure cylinder. 
It is important that the copper tubes of the intercooler be kept 
clean. As the oil carried over by the air tends to deposit on the 
tubes, they should be so arranged as to be readily accessible for 
cleaning. The intercooler of the Schram (English) two-stage com- 
pressor is a vertical chamber, also filled with small tubing. The 
water enters at the bottom, passes up through one-half of the tubes 
and down through the other half, the lower wateryhead being di- 
vided accordingly. The air from the low-pressure cylinder enters 
at the top of the intercooler, passing out at the bottom into the 
high-pressure cylinder. 

Although the relatively small intercoolers of ordinary two-stage 
compressors are imperfect in their action, as has been pointed out, 
it is nevertheless possible to attain a high degree of efficiency from 
intercoolers of large capacity. A well-known example may be 
cited: the plant of the Paris Pneumatic Supply Co., in which 
Riedler two-stage compressors are used. Spray injection is applied 
to both cylinders, and also a plain intermediate receiver of very 
large capacity, but without tubes. The air is compressed to 88 
pounds, and the indicator diagrams of the air cylinders exceed in 
area the true isothermal diagram by only 12.07 per cent.* That 
is, the work done twice is about 12 per cent, of the total work, the 
total efficiency having the high value of 77 per cent. 

To show the results obtained by thorough cooling of the air 
between the cylinders, a comparison of the work done by single- 
and double-stage compression may be made. Frictional losses 
will be omitted in each case, and no account will be taken of the 
cooling due to the cylinder water-jackets. 

I. A single-stage compressor, producing a gauge pressure of 

''^Proceedings Institution of Civil Engineers, London, Vol. CV, p. 180. 



no COMPRESSED AIR PLANT 

70 pounds at sea-level, with a 24-inch cylinder and a piston speed 
of 400 feet per minute, will have a capacity in terms of free air at 
normal temperature of 1,256 cubic feet per minute. For adiabatic 
compression, the mean cylinder pressure will be 33.83 pounds and 
the horse-power 184.38. 

2. For doing the same work in a two-stage compressor, provided 
with an intercooler capable of reducing the temperature of the air 
to the normal between the cylinders, it may be assumed that the 
low-pressure or intake cylinder has the same diameter, 24 inches, 
and that the pressure produced in it is 35 pounds. The mean press- 
ure (adiabatic), corresponding to 35 pounds terminal pressure, is 
21.6 pounds, and the horse-power 11 8. 19. The diameter of the 
high-pressure cylinder, under the assumed conditions, is found by 
making the piston area inversely proportional to the increase in 
absolute pressure of the air delivered to it by the low-pressure 
cylinder, i.e., in the ratio of 14.7 : 35 + 14.7 = i : 3.38. This gives 
an area of 135 square inches, equivalent to 13 inches diameter. 
Compressing in this cylinder from 35 to 70 pounds gauge, the mean 
effective pressure will be 28.74 pounds, and the horse-power, 46; 
or a total for both cylinders of 11 8. 19 -f 46 = 164.19 horse-power. 

Compared with the power required for doing the same work 
in the single cylinder, this shows a saving of: 184.38 — 164.19 = 
20.19 horse-power, or about eleven per cent. The theoretically 
perfect cooling between the cylinders here assumed would not be 
attained in- ordinary practice, however, and the frictional loss in 
the stage comxpressor would probably be a little greater than in 
the single-cylinder machine; so that the net gain due to inter- 
cooling may in this case be taken at, say, seven to eight per cent. 
The saving is considerably increased in dealing with Jiigher 
pressures. (For " Stage Compression at High Altitudes," see p. 219.) 

The advance made in recent years in the design of intercoolers 
is further illustrated by Fig. 56, showing a new design of the 
Ingersoll-Rand Co. It is provided with pipe connections for drain- 
ing off the water deposited as a result of the reduction in tempera- 
ture. These coolers may be employed also as " receiver-after- 
coolers," which are now considered as almost essential adjuncts 



COMPOUND OR STAGE COMPRESSORS 



III 



of well-installed large plants. (See Chapter XI.) A similar 
appliance may be employed advantageously as an ante-cooler for 
the intake air. 

The useful effect of small intercoolers, such as are frequently 
mounted above the cylinders of straight-line compressors, should 




Fig. 56. — Vertical Intercooler. Ingersoll-Rand Co. 



not be misunderstood nor exaggerated. It must be remembered 
that the best economy in air compression is obtained only by 
cooling during compression and before the air leaves the cylinder. 
Hence, in addition to the intercooler, the largest possible water- 
jacket area should be provided. 

The relation between the compression curves of a two-stage 



112 



COMPRESSED AIR PLANT 



compressor is shown in Fig. 57, the adiabatic and isothermal curves 
being also laid down.* These cards, not accurately reproduced 
here, were taken from a pair of cylinders measuring 7 J and 14 X 16 
inches, compressing to no pounds gauge, at 135 revolutions per 
minute, or 360 feet piston speed. Initial temperature of cooling 
water, 55°; temperature at discharge from jackets and intercooler, 




Fig. 57. — Combined Air Card of Two-Stage Compressor. 



62° F. Several points are to be noted in connection with these 
combined two-stage cards : 

First. The overlapping of the high- and low-pressure cards in- 
dicates a loss, because the work represented by the area of the over- 
lap is in reality work done twice. This is the result of the drop 
in pressure between the cylinders, which is caused by the resistance 
presented by the discharge valves of the low-pressure and the inlet 
valves of the high-pressure cylinder, together with the friction 
in the air passages and intercooler. While this loss is unavoidable, 

* This combined indicator card, which does not show all the minor irregularities 
in the lines, is from a Rand cross-compound compressor. It accompanies an ar- 
ticle by F. A. Halsey, on "The Analysis of Air Compressor Indicator Diagrams," 
American Machinist, IS/Larch 3d, 1898, p. 158, and is reproduced here by permission. 



COMPOUND OR STAGE COMPRESSORS II3 

it should be reduced as much as possible by making the valves, 
ports, and connecting passages of ample size. 

Second. As with single-cylinder dry compressors, the com- 
pression lines of the individual cylinders of most stage compressors 
depart but little from the adiabatic curve. Aside from the thermo- 
dynamic advantage of dividing the total compression between two 
or more cylinders, and thereby lowering the average and final tem- 
peratures, it is the intercooler that must be relied on for furnishing 
the chief element in economical working. By its abstraction of heat 
the volume of air entering the second cylinder is reduced, so that 
-pyn = 1.4 ^ constant becomes approximately P V = C, on beginning 
the second stage of compression. But the compression line again 
rises rapidly from this point and continues not far below the 
adiabatic. 

Indicator cards from dry compressors which do not show 
approximately this relation between the lines are always open to 
suspicion. A leaky piston, for example, will lower the compression 
curve and make it appear that much better work is being done than 
is really the case. It may be observed that, other things being 
equal, a lower curve is often obtainable from a small than from a 
large compressor, because the ratio of area of water-jacket to the 
volume of the cylinder is greater. 

In constructing and reading a combined indicator card from 
both cylinders of a stage compressor (like that shown in Fig. 57), 
the adiabatic line applying to the compression in the second cylinder 
should be represented in its proper place. The complete graphic 
relation between the several heat curves is thus set forth. 

Third. It is an advantage of stage compression that there is 
practically but one clearance space — that in the low-pressure cylin- 
der — and, as the air in this cylinder is at a low pressure, the re- 
sulting reduction in net volumetric capacity is moderate, for it is 
evident that the loss due to clearance is proportionately less for low 
than for high pressures. The piston clearance of the high-pressure 
cylinder cannot affect the volume of air delivered, because all the 
air discharged from the low-pressure cylinder goes to the high- 
pressure and, barring leakage, must pass through it. 
8 



114 . COMPRESSED AIR PLANT 

The heating of the cyHnder walls and pistons reduces somewhat 
the working volumetric capacity of an air compressor because, as 
the entering air is warmed, a smaller weight of it is taken into 
the cylinder at each stroke. Although the degree of this heating 
cannot be formulated, it is obvious that it is less in a two-stage 
than in a single-cylinder compressor; for, aside from the effect of 
the intercooler, the smaller quantity of heat developed in each 
cylinder is more efficiently dealt with by their respective water- 
jackets. 



CHAPTER VII 

AIR INLET VALVES* 

The proper design and working of the inlet or suction valves 
exert an important influence on the efiiciency of the compressor, 
and perhaps no other one portion of air-compressor mechanism has 
received so much attention. Nevertheless, that there are still 
wide differences of opinion as to the best design for inlet valves 
is evidenced by the great variety of types used by compressor- 
builders and the lack of clearly defined distinctions as to their 
applicability under different working conditions. Reference to 
almost any compressor catalogue will show that the purchaser has 
a choice of several types, with but little to guide him in making a 
selection. 

In the older forms of wet compressor various patterns of clack- 
valve were employed, as exemplified in the Dubois-Frangois com- 
pressor. Though not now used in this country, they have by no 
means been abandoned in Europe; witness the Guttermuth valve 
and the elaborate, cam-controlled clack-valves of some large com- 
pressors built by Schneider & Co., Creusot, France. For years 
poppet valves of numerous types held the field in the United States 
almost exclusively. They are furnished with springs, and are usu- 
ally actuated solely by difference of air pressure ; though in a few 
designs mechanically controlled poppets were introduced, such as 
those of the old Rand mechanical valve-gear and others, examples 
of which are still occasionally to be found in use. While poppet 
valves have continued in favor for certain kinds of service, and are 
likely to remain so, many other forms of inlet valve have been 

* This chapter is devoted chiefly to spring poppet valves and others which 
operate by difl'erence of air pressure. For discussion of those inlet valves whose 
movements are under mechanical control, see Chapter IX. 

115 



Il6 COMPRESSED AIR PLANT 

successfully applied in the course of the development of the 
modern compressor. Modifications of the Corliss rotary steam 
valve, first used in the Norwalk compressor, have now been 
adopted in compressors of many other makes, such as the Nord- 
berg, Sullivan, Laidlaw-Dunn-Gordon, and Alhs-Chalmers. There 
are at least two inlet valves which cannot be included in any of the 
other classes, viz. : the Sturgeon valve, placed in the cylinder head 
and operated by frictional contact with the piston rod, and the 
ingenious Ingersoll- Sergeant piston inlet, which opens and closes 
by its own inertia at the end of each stroke. Both of these operate 
under fixed conditions, independently of differences in air pressure 
within and without the cylinder. 

The two chief requisites of all inlet valves are: i. That they 
shall have a sufficient area of opening to permit free entrance of 
the air. 2. That they shall open readily near the beginning of the 
stroke, with a minimum of resistance, remain open until the end 
of the stroke, and then close promptly. 

There are several questions affecting the design and operation 
of the usual types of inlet valve, which are closely related to the 
working of the air cylinder itself. The point of the stroke at which 
the inlet opens should depend on the piston clearance and the 
air pressure under which the compressor is working. Spring-con- 
trolled valves, or those operated mechanically, are sometimes in- 
correctly designed or set, so as to open exactly at the beginning 
of the stroke or a fraction later; in which case the clearance air 
is first exhausted through the valves and then, as the piston ad- 
vances, the outside air begins to enter. This being so, it is evident 
that no clearance at all would be shown on the indicator card. 

As already pointed out, although piston clearance causes a 
reduction in volumetric capacity of the cylinder, it not only does 
not involve a corresponding loss of work, but is in reality beneficial, 
in assisting to overcome the inertia of the reciprocating parts of the 
compressor. A large part of the work expended in compressing 
the clearance air is thus recovered. But when the clearance air is 
exhausted wholly or in part by a premature opening of the inlet 
valves, the work represented by it is lost. With spring-controlled 



AIR INLET VALVES II7 

poppet valves the proper adjustment is a question of the strength 
of the spring, and since the effect of clearance varies with the air 
pressure, the valves must be regulated for the pressure carried in 
each particular case. Any exhaust through the inlet valves is 
readily detected by the noise. When they are properly set, the com- 
pressor works more smoothly and the power consumed is slightly 
reduced. On the other hand, if the valves open too late in the 
stroke — due, for example, to a temporary reduction in working 
pressure — a little more power is required, this condition being 
shown by the slight drop in the re-expansion line at the point b 
(Figs. 48 and 57). 

For inlet valves which are opened and closed mechanically, 
an adjustment to the working conditions is even more imperative 
than in the case of valves controlled only by springs. If incorrectly 
set or timed with respect to the stroke of the piston, they may be 
forcibly opened too early in the stroke or closed before the end. 
Premature closing obviously reduces the volume of intake air, and 
with it the volumetric capacity of the compressor. Its effect on the 
indicator card is to lower the compression line near the beginning 
of the stroke, so as to approach the isothermal curve and make 
it appear that the compressor is doing abnormally good work. 

The total area of the inlet ports varies greatly in compressors of 
different makers. It is sometimes as small as 4 or 5 per cent, of 
the piston area, running up to a probable maximum of 12 to 15 
per cent. As the proper area is really a function of the piston 
speed, it may be made less for slow- than for high-speed compres- 
sors. However, in one of the Leyner 2-stage compressors, with a 
22-inch low-pressure cylinder and running at the moderate piston 
speed of 390 feet, the intake port area is 14.2 per cent, of the piston 
area. (The Leyner valves are of a special type, described hereafter.) 
To insure freedom from excessive frictional resistance against the 
inflow of air, the inlet area, under average conditions and for ordi- 
nary forms of valve, should be not less than, say, 8 or 10 per cent, 
of the piston area. But extremes should be avoided. If poppet 
valves are made unnecessarily large, their inertia becomes great; 
and if too numerous, there are not only more parts to care 



ii8 



COMPRESSED AIR PLANT 



for, but valuable water-jacket area on the cylinder heads must be 
sacrificed. 

In the two-stage, straight-line, "Hurricane-inlet" compressors, 
of the Ingersoll-Rand Co., type AA-2, for cylinders from 15 -in. to 
24-in. diameter, the inlet area of the low-pressure or intake cylin- 
der averages 13.2 per cent, of the piston area. For the high press- 
ure cylinders of the same compressors, the poppet valves have an 
average inlet area of about 1 1 per cent. The inlet area of the du- 
plex, two-stage compressors, type O-2 of the same builders, aver- 
ages 13.6 per cent, of the piston area for both low- and high-press- 




FiG. 58. — Norwalk Poppet Inlet Valve. 

ure cylinders. In the two-stage compressors of the Laidlaw- 
Dunn-Gordon Co. the percentage is from 12 to 14. 

Poppet Inlet Valves. One of the commonest forms is the mush- 
room valve, two types of which are shown in Figs. 58 and 59. 
While the total inlet area should be ample, there are two special 
requirements in the case of ordinary poppet valves: (i) the area of 
each individual valve must be moderate, or the valve will become 



AIR INLET VALVES 



119 



too hea\y, causing unnecessary injury to the valve seat, and by 
its inertia too great a resistance to the control of the spring; 
(2) the lift must be small, in order to attain prompt opening and 




Fig. 59. — Laidlaw-Dunn-Gordon Poppet Inlet Valve, 

closure, and to reduce '^ chattering," as well as wear. For these 
reasons the total area required is furnished by a number of in- 
dependent valves, generally from four to six, which are set m each 
cylinder head. 



I20 COMPRESSED AIR PLANT 

The valve is of steel or bronze, with an easily removable bronze 
seat, the contact surfaces being ground true and the seating often 
coned. To control and close the valve promptly its stem is pro- 
vided with a spiral spring. The stem w^orks in guides, forming 
part of the seat and valve casing, which is screwed into the cylinder 
head so as to be readily removed when necessary for adjustment 
or repairs. Brass springs are used, to avoid the effects of corrosion, 
and must be easily compressible to allow the valve to open freely 
under a small difference of pressure; that is, as early in the stroke 
as possible after the clearance air has re-expanded. The springs 
should be made of the best material and accurately proportioned to 
present no more than the minimum requisite resistance to opening. , 
Under a ctual working conditions the pressure of the springs varies 
from, say, three ounces to eight or even ten ounces per square inch 
of valve area. 

Ordinary poppet valves are opened by the atmospheric pressure 
from without, when a certain degree of rarefaction of the air inside 
the cylinder has been produced by the movement of the piston; 
in other words, when the difference of pressure, after the clearance 
air has re-expanded, becomes sufficient to overcome the resistance 
of the spring, and compress it. In accomplishing this the piston 
must advance some distance before any air can enter the cylinder. 
The loss of volumetric capacity thus caused, in terms of free air, is 
probably rarely less than two to three per cent., and is often more. 
At sea-level a spring pressure of five ounces per square inch of 
valve area causes a loss of about two per cent. The diagram. 
Fig. 60,* shows the effect of spring resistance in reducing the volu- 
metric capacity of a compressor at different altitudes, from sea- 
level to 1 5,000 feet elevation. 

With spring-controlled poppets there is more or less irregu- 
larity in the entrance of the air, because, while the pressure of the 
outside air tries to open the valve, the action of the spring tends to 
keep it closed. This often produces "chattering" or '^ dancing" of 
the valves, and has led among other things to the introduction of 
various mechanical devices for definitely controlling them, as will 

* Reproduced by permission from Engineering News, May 30th, 1901, p. 391. 



AIR INLET VALVES 



121 



be noted later. As the springs lose their original elasticity, and 
undergo alterations in strength, they require regulation from time 
to time; outside adjusting nuts on the valve stems may be provided 













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122 



COMPRESSED AIR PLANT 



for this purpose. If the springs be too slack, the chattering increases ; 
if too tight, the valves will open late in the stroke, and the intake air 
occupying the cylinder will have a density less than that of the 
atmosphere. But, aside from spring resistance, the rate of inflow 
of the intake air is variable. This is due to the variation in speed 
of the piston. When its speed is greatest, at the middle of the 



Deliv .Press, =65 Lbs. 




Fig. 6i. 



stroke, the rate of inflow of air is at the maximum. While the pis- 
ton is moving slowly, near the beginning and end of each stroke as 
the crank turns its centers, the relatively small negative pressure 
becomes insufficient to open the valves and keep them open against 
the strength of the springs. The effective length of stroke is thus 
shortened. 

The total valve resistance, including that due to throttling of 
the intake air and friction in passing through the ports, must be 
kept as small as practicable, but can never be entirely eliminated. 
With some forms of inlet valves, other than spring poppets, the re- 
sistance becomes very small, and sometimes almost inappreciable. 
Its usual effect is shown on the diagram. Fig. 6i. There is generally 
sufficient resistance to keep the admission line, A C, at an apprecia- 
ble distance below the atmospheric line, D E, throughout the 
stroke; the amount of loss from this cause being measured by the 



AIR INLET VALVES 1 23 

area of the indicator diagram lying below the atmospheric line. If 
the inlet area be too small or the valves poorly designed, the result- 
ing negative pressure may amount to one or two pounds per square 
inch. The point B, where the compression line crosses the atmos- 
pheric line, is the point of the stroke which must be reached by the 
piston before any useful work is done, and the volume passed 
through in travelling from A to B represents the loss in volumetric 
capacity from this cause. The total loss of volumetric capacity, 
including that due to piston clearance, is represented by the length 
of A B + C E, and the volumetric efficiency of the compressor is 
measured by the length of the Hne B C, projected on the atmos- 
pheric line. 

Notwithstanding certain inherent disadvantages, the poppet 
valve in different forms is widely used, for both inlet and discharge. 
It is simple in construction, easily regulated, and in case of leak- 
age, due to cutting and unequal wear of the seating surfaces, is 
readily removed and re-ground. In stage compressors it is often 
used for the high-pressure cylinder, even when some other type is 
preferred for the low-pressure. Poppet inlet valves not infre- 
quently cause trouble by sticking in their seats on account of the 
accumulation of gummy oil. Or, they are sometimes clogged by 
deposit of carbonaceous matter from decomposition of the lubri- 
cant, produced by excessive heating of the cylinder. The valves 
should be kept clean, and are therefore designed to permit ready 
access. 

One of the recent forms of Norwalk two-stage compressor 
has a special poppet inlet valve, designed for use when it is desired 
to employ air at two different pressures, obtained from a single com- 
pressor. In stage compression, though the air is actually pro- 
duced at two pressures, of, say, 25 to 3c and 80 to 100 pounds, re- 
spectively in the lov/- and high-pressure cylinders, yet, if a part 
of the volume delivered by the low-pressure cylinder be drawn 
from the intercooler, the high-pressure cylinder fails to work satis- 
factorily. The air remaining in the intercooler expands to a lower 
pressure before going to the high-pressure cylinder, so that the ratio 
of compression in this cylinder is increased, and the heat gener- 



124 



COMPRESSED AIR PLANT 



ated is raised to a correspondingly higher degree. With such a rise 
in temperature as would be produced by increasing the ratio of 




> 

I 



compression from, say, three to fifteen or twenty, proper lubrica- 
tion is impossible, and the conditions would be favorable for an 
explosion in the cylinder. 



126 COMPRESSED AIR PLANT 

This difficulty is met by using '' skip- valves " (Fig. 62) as inlet 
valves of the high-pressure cylinder. They are designed to open, 
and remain open, whenever the high -pressure inlet air falls below 
the normal, by reason of having drawn off a portion of the air from 
the intercooler. The high-pressure cylinder is thus temporarily 
unloaded in part, since the air entering at each stroke is returned 
to the intercooler. The skip-valve is a mushroom spring-poppet, 
D E, carried in the guides A, A. Above the valve is a small, spring- 
controlled plunger B, the space below which is occupied by air at 
intercooler pressure. When this pressure falls below that for 
which the spring C is set, the plunger advances and forces open 
the inlet valve, holding it open until the intercooler pressure rises 
sufficiently to cause the plunger to recede. The valve is then free 
to work automatically in the usual manner. The action of the 
valve thus adjusts itself constantly to the varying pressure of the 
intake air coming from the intercooler; and the variation in con- 
sumption of power by the high-pressure cylinder is taken care of 
by the governor applied to the steam end of the compressor. 

Ingersoll-Rand " Hurricane Inlet " Valve. In many of the 
IngersoU-Rand compressors, the inlet valves are placed in the 
piston. Fig. 63 shows a longitudinal section through the cylinder 
and piston; Fig. 64, an enlarged section of the piston and valves. 
The piston is hollow and into its rear end is screwed a hollow 
back piston-rod, for admitting the air. The pipe is provided with 
an ordinary stuffing box in the cylinder head. There are two 
large, ring-shaped valves (one in each face of the piston), of T 
cross-section and made of oil-treated, annealed steel. The valves 
are a little smaller in diameter than the piston and are held in 
place, without springs or other connection, by a guide-plate bolted 
to the piston face. Their play is limited by these guide plates 
(see Fig. 64), which contain a series of circular ports, furnishing 
additional area for the passage of the air. The valves are read- 
ily taken out for regrinding when necessary, by removing the guide 
plates which are attached to the piston by tap-bolts. 

While the compressor is running, the air is drawn in through 
the hollow piston-rod in practically a constant stream, passing 



AIR INLET VALVES 



127 



through either valve, first into one end of the cyhnder and then 
into the other. At the beginning of each stroke the valves are 
alternately opened and closed by their own inertia, as the piston 
reverses its motion. The valve in that face of the piston which is 
toward the direction of movement is always closed, while the other 
is open for the passage of the air entering through the hollow rod 
into the cylinder behind the piston. On account of the large di- 




FiG. 64. — Ingersoll-Rand "Hurricane Inlet, " Enlarged Section. 



ameter of the valves their throw, or lift, is small — in ordinary com- 
pressors say three-eighths inch. 

In the "Hurricane-Inlet" cylinders the area of the hollow pis- 
ton rod, or inlet tube, is usually 13 to 14 per cent, of the piston 
area. Though the actual port area of the valve is less than this 
— say 8 per cent. — the velocity of flow is moderate. The net port 
area afforded by this valve may be less than in some compressors 
having a group of inlet valves, but is found to be sufikient, be- 
cause the inlet is concentrated in a single opening. It is probable 
that during admission there is less difference between the pressure 
of the air taken into the cylinder and the atmospheric pressure 
than with any form of spring-controlled valve; for, meeting with 



128 COMPRESSED AIR PLANT 

no resistance due to springs, the air enters freely. Moreover, 
when the end of the stroke is reached, the inflow of air is momen- 
tarily checked, while the piston is reversing, and the momentum 
of the column of air in the inlet tube tends to cause a slight in- 
crease in the density, and therefore in the weight, of the body of 
air already in the cylinder. 

These valves wear well, and their use permits a piston speed 
high enough to conform with modern high-speed compressor 
practice. Other advantages are: the cylinder castings are sim- 
plified; the space in each cylinder head that would otherwise be 
occupied by inlet valves is utilized for additional water-jacket 
area; and the number of moving and wearing parts is reduced. 
It has been objected that, since the hollow rod and piston are 
necessarily heated, these advantages are partly offset by a rise in 
the temperature of the intake air, in its passage into the cylinder; 
and that therefore the weight of air in the cylinder is relatively 
less than if it had entered by a more direct path. But it does not 
appear that this objection is well founded. The air enters the 
hollow piston rod in a single large volume, instead of being di- 
vided into comparatively small areas of flow. It has little oppor- 
tunity to absorb heat until it reaches the valve, because only a 
thin film of the rapidly moving air in the inlet tube is in contact 
with the tube itself. 

No positively conclusive tests have yet been made, as to the 
relative heating of the intake air in " Plurricane-Inlet " and in pop- 
pet-valve compressors. Some heat is undoubtedly absorbed by 
the air in passing in the thin sheet through the valve port in the 
piston face; but thermometric observations, taken inside the in- 
let tube and piston, at speeds of 40 and 120 revolutions per minute, 
show an increase of temperature of not over 5° Fah. at the lower 
speed and even less at the higher speed. It seems unlikely that 
any better results are obtainable from either poppet or Corliss 
inlet valves. 

The "Hurricane-Inlet" is a modification of the well-known 
Ingersoll- Sergeant "Piston-Inlet" Valve, employed for many 
years in the compressors of these builders. The older type is 



AIR INLET VALVES 



129 



somewhat heavier in proportion to the area of opening and is 
quite different in design. Like the "Hurricane-Inlet," it rests in 
the seat without springs or other connections; but in the piston 
casting there is inserted a series of small studs, which pass through 
slots in the ring valve and so limit the throw. ' 

Johnson Valve. In the Johnson compressor, built in England, 
there is a single poppet inlet valve of the gridiron type at each 






PLAN OF LIFTER 



Fig. 65. Fig. 66. 

Figs. 65 and 66. — Johnson Air Valves. 



end of the cylinder. It has a large area, with a small lift, and is 
mounted in a peculiar way on the same spindle with the discharge 
valve (Figs. 65 and 66) . Both valves are rendered easily accessible 
by being placed in a chamber projecting horizontally from the end 
of the cylinder. This chamber is closed by a cast-iron plate held 
in place by a yoke and set-screw. The lift of the valve is con- 
trolled by an outside adjusting nut, c, on the spindle. The inlet 
valve is provided with a ''lifter" (Fig. 65, d) by which it can be 
raised trom its seat and thrown out of use, if it be desired tempora- 
9 



I30 



COMPRESSED AIR PLANT 



rily to make the compressor single-acting. The Johnson valve 
closes by gravity only, no springs being used. 

Humboldt Rubber Ring Valve. The older form of Humboldt 
wet compressor (see Fig. 44) has a simple and ingenious valve 
(Fig. 6j). It consists merely of a rubber ring of round cross- 
section v^hich covers a series of horizontal slots, or ports, in a cyl- 
indrical casting set in the top of each air chamber. Three of these 
rings, a, v^ith the slots, /, comprise the inlet valves in each end of 
the air cylinder; the casting, t, in which they are placed forming 
a part of the valve- chamber cover. The casting, c, is strengthened 




Fig. 67. — Humboldt Rubber Ring Valve. 

against the internal pressure by a series of webs, d. As the 
pressure in the air chamber falls during the inlet stroke the at- 
mospheric pressure expands the rubber rings, forcing them away 
from the slots, and allowing air to enter. Then on the reversal of 
the stroke the elasticity of the rings causes them to tighten up on 
their seats and close the ports. The valve openings are relatively 
large and permit free entrance of air. The discharge valve, &, has 



AIR INLET VALVES I3I 

the same construction, but consists of a single ring only, of larger 
diameter and cross-section. These rubber valves are found to last 
well, as they are kept wet and are not exposed to any great degree 
of heat. They would be entirely unsuitable for dry compressors. 
Similar rubber valves are used in a wet compressor built by the 
Dingier ^lachine Works, Zw^eibruecken, Germany. 

The Guttermuth Valve is used in a later form of compressor built 
by the Humboldt Machine Works. It is a spring clack-valve, 
made of a rectangular plate of thin steel and provided with a grid 
seat. One side of the plate is coiled in a spiral, through the center 
of which passes a stationary rod or spindle, the inner edge of the 
spiral being inserted in a longitudinal groove in this spindle. By 
placing several valves side by side any desired area of opening can 
be furnished. To avoid the harmful effects' of inertia, the valves 
are made of extremely thin plate, with delicately adjusted and 
sensitive springs, and by so arranging them that the current of 
air in passing through the valve into the cylinder undergoes but 
slight changes of direction, any serious eddying of the air around 
the edges of the plate is prevented. 

Leyner Flat Annular Valve. This recent form of valve, to- 
gether with its arrangement on the cylinder heads, is shown by Figs. 
68 and 69. Fig. 68 comprises a longitudinal section through the 
adjacent ends of the low- and high-pressure cylinders of a straight- 
line, two-stage compressor, indicating incidentally the circulation 
of the air through the intercooling tubes of both cylinders, as de- 
scribed in Chapter VI. At each end of the cut, left and right, is 
an outline cross-section, respectively of the low-pressure and 
high-pressure cylinder heads, showing the groups of intercooler 
tubes, with the valves themselves and their ports. 

The inlet and discharge valves being similar in form, a descrip- 
tion of the inlet only will be given (Fig. 69). It consists of a thin 
steel plate cut in a peculiar form. The outer, or seating portion, 
is a narrow annulus, with two slender internal arc-shaped strips 
terminating in a central ring, which is locked against the cylinder 
head by a steel nut encircli-ng the piston-rod, thus holding the 
valve in place. The arc-shaped strips, connecting the seating part 



132 



COMPRESSED AIR PJ ANT 




> 



S 

3 



AIR INLET VALVES 



133 



of the valve with the fixed central ring, are sufficiently long and 
flexible to serve as springs, and to permit the valve to open and 
close freely under very small differences of pressure. There is but 
one inlet and one delivery valve at each end of the cylinder. The 
inlet ports, D, D (Fig. 68), four in number in each cylinder head, 
are curved, slot-like openings, arranged in the form of a circle. 
There are six similar but smaller discharge ports, E, E. Total area 
of inlet ports is about fourteen per cent., and of discharge ports, 
nearly nine per cent, of the piston area. The discharge valves are 
held in position by the hollow conical casting surrounding the 




Fig. 69. — Leyner Annular Inlet-Valve 

piston-rod stuffing-box. Their height of lift is limited by the 
stops, shown near F, F. 

This valve is simple in design, without separate springs, and con- 
sists of one part only. It cannot be doubted that the resistance 
to opening of the inlet valves is extremely small. As the clearance 



134 COMPRESSED AIR PLANT 

volume in these compressors is small (1.02 per cent, of the cylinder 
volume in the 14-inch high-pressure cylinder mentioned in the 
foot-note*), a high volumetric efficiency is stated to have been ob- 
tained, a number of tests showing it to range from 94.6 to 97 per 
cent. 

Arrangements for Admitting Inlet Air to the Compressor. It is 
of great importance that the intake air shall be as cool as possible. 
The colder the air the smaller is the volume occupied by a given 
weight of air taken into the compressor cylinder, and the greater the 
output. Taking in warm air involves loss of capacity and of econ- 
omy in production. Mr. Frank Richards points this out in a con- 
vincing and simple way.f ''The volume of air at common tem- 
peratures varies directly as the absolute temperature. With the 
air supply at 60° its absolute temperature is 521°, and its volume will 
increase or decrease -g-iy for each degree of rise or fall of tempera-' 
ture. Therefore, if in securing the supply of air we can get a dif- 
ference in our favor of 5° ... we accomplish a saving of about 
one per cent. If a difference of temperature of 10° can be secured 
two per cent, is saved," practically without cost. The practice of 
taking air from the engine-room is a common one at mines, and is 
bad not only because such air is usually heated to a considerable 
degree, but is apt also to be charged with dust which causes un- 
necessary wear of valves and piston. 

Some means should be provided to convey to the com.pressor 
fresh air, taken preferably from some point outside of the building. 
A box or pipe of wood is better than one of iron, because of the 
smaller conductivity of wood. Its cross-section should be suf- 
ficient, say, at least one-half the area of the cylinder, to avoid loss 
from friction. To make such a connection conveniently the inlet 
valves should be enclosed in an external air chest on each end of 
the cylinder. Compressors having a single inlet valve, such as 

* In a communication to the author the makers state that repeated tests of a 
14 and 22 X 26 inch, 2 -stage compressor show a loss of intake pressure of only 
0.9 ounce. On a card made with a 20-scale spring, this would be represented by a 
difference of the inappreciable amount of 0.003 inch, between the intake and at- 
mospheric lines. The frictional loss through the delivery ports of the same com- 
pressor is 3 ounces. 

t " Compressed Air," p. 55. 



AIR INLET VALVES 1 35 

the Norwalk, Ingersoll-Sergeant, Sturgeon, etc., are better adapted 
than some of the others for making this arrangement. In any 
case, care should be taken to prevent the entrance of dust, leaves, or 
rubbish. If the inlet be left open, particles floating in the air may 
be drawn in by the strong current, and obstruct the valves or in- 
jure their seats and the smooth working surfaces of piston and 
cylinder. In such a design as that of the Ingersoll-Sergeant piston 
inlet, it is essential that the outer end of the hollow rod be covered, 
because in case of derangement the valves are not so accessible as 
ordinary poppets. This protection is provided in recent designs 
of this compressor. By building a suitable conduit from the out- 
side of the compressor house to the air box enclosing the inlet valves, 
it is obvious that a greater saving can be effected in winter than in 
summer, but even in warm wxather some advantage is gained, 
especially if the conduit opens on the north side of the building, out 
of reach of the sun's direct rays, and is carried vertically to some 
height above the ground level. 



CHAPTER VIII 

DISCHARGE OR DELIVERY VALVES 

The conditions affecting the action of the discharge valves of a 
compressor are wholly different from those which govern the suction, 
or inlet valves. While the latter must be capable of opening 
under very small differences of pressure, the discharge valves are 
subjected to a heavy pressure on both sides. Furthermore, owing 
to unavoidable irregularities in the use of the air, the receiver 
pressure usually fluctuates considerably, so that the point of the 
stroke at which the discharge valves open cannot depend solely on 
the conditions, as to the ratio of compression, etc., under which 
the compressor itself is working. The time of opening must de- 
pend also on the relation between the variable pressures in 
cylinder and receiver. 

For this reason, the sensitiveness of operation essential in inlet 
valves is unnecessary for the discharge valves. The chief require- 
ments are that they shall be free to open when the cylinder pressure 
exceeds that of the receiver, shall fit accurately on their seats, and 
close promptly at the end of the stroke. Delay in closing, or leak- 
age between valve and seat, are far more serious than in the case of 
inlet valves, because these defects are equivalent to an increase of 
the piston clearance and consequent reduction of the volumetric 
capacity of the cylinder. The leakage of even a small quantity of 
compressed air back into the cylinder is equivalent to the loss 
caused by an abnormally large clearance space. The conditions 
under which discharge valves operate, therefore, are such as to 
afford a relatively limited field for innovation or improvement, 
as compared with inlet valves. 

Poppet Discharge Valves. Aside from a few designs in which 

136 



DISCHARGE OR DELIVERY VALVES 



137 



mechanical control in some form is introduced (see Chapter IX), 

nearly all discharge valves are of the poppet type. They are made 

heavier than inlet valves, with stronger springs to reduce hammer- 

j ing on their seats. Though varying in details of construction, they 




Fig. 70. — Laidlaw-Dunn-Gordon Poppet Discharge Valve. 

may be represented fairly by the accompanying figures. Several 
other designs are also shown in the various sections of air cylinders 
illustrated in the preceding pages. Two of the ordinary forms of 
cup-shaped poppet, with internal springs, are shown in Figs. 70 
and 71. Occasionally they are of the mushroom type, somewhat 



138 



COMPRESSED AIR PLANT 



similar in shape to the inlet valve (Fig. 58), the spring then encir- 
cling the spindle. The valve may be of steel or bronze, with a 
bronze seat. To make it easier to keep them tight, the seating 
surfaces are usually coned. A group of several poppet valves are 
commonly employed, in order to avoid making them of large size 
and weight. The inertia of heavy valves causes destructive wear, 
under their high working pressure. Each valve should be readily 
accessible for adjustment, re-grinding, or renewal. They are there- 




FiG. 71. — Norwalk Poppet Discharge Valve. 



fore covered by caps screwed into the outer cylinder head; or, in 
some makes., by plates bolted on over the valve chamber. 

Cataract-Controlled Poppets. In another type of poppet dis- 
charge, the valve is not only provided with a spring, but its action 
is further modified by attaching the valve stem to the piston of a 
small cataract cylinder, containing either air or oil. This is to 
ease their movements and avoid hurtful shocks.* Oil-cataract 

* Similarly controlled poppets are also employed as inlet valves by some Euro- 
pean compressor-builders. 



DISCHARGE OR DELIVERY VALVES 



139 



valves are used, for example, in the compressors built by Schuech- 
termann and Kremer, Dortmund, Germany;* air-cataracts in 
those of R. Meyer, Muhlheim-Ruhr; G. A. Schuetz, Wurzen; 
Menck and Hambrock, Altona, and the Humboldt Machine Works, 
Kalk. (The rubber ring discharge valve, of the last-named build- 
ers, has already been referred to, in connection with Fig. 67.) 

These valves are employed to a considerable extent in Europe, 
but are not well known in this country. Some of them are 




Fig. 72. — "Express" Poppet Valve, Riedler Compressor. 

very^ satisfactory, provided the piston speed be slow; for high- 
speed compressors they do not work with sufficient promptness 
to prevent " slip" or leakage of some of the compressed air back 
into the cylinder. The chief object sought in these cataract move- 
ments is attained in another way — by the partial control of an ac- 
companying Corliss valve — in the " Cincinnati " valve gear of the 
Laidlaw-Dunn-Gordon Co., described in Chapter IX (see Fig. 76). 
Riedler Discharge Valve. A poppet discharge valve entirely 
different in design is shown in Fig. 72, representing one of several 
patterns employed in the Riedler compressors. This is a light, 

* Described in London Engineering, Dec. i2lh, 1902. 



I40 COMPRESSED AIR PLANT 

cylindrical valve, A, provided with packing rings D. The cylinder 
in this case is vertical, and the piston, L, carries at its periphery the 
plate P, held in place by the stud N and the spring M. When 
closed the valve seats on the plate E, being held against it by the 
air pressure in the discharge passage acting on the under side of the 
upper flared end. In this position the round air ports near the 
lower edge of the valve are closed by the valve guide, at C,C. As 
the piston advances, and when the cylinder pressure exceeds that in 
the receiver, the valve is opened by the air pressure on the upper 
side of the flared end. This movement of the valve is cushioned 
by the air trapped above the guide, B, B. On reaching the end of 
the stroke, the plate P, on the piston, strikes the lower edge of the 
valve and closes it against its seat E, the shock being cushioned by 
the springs F and M. The double cushioning, in both opening 
and closing, tends to durability; and, moreover, it should be re- 
membered that, when the plate P strikes the valve, the crank is 
nearly on its center, so that the piston is moving very slowly. The 
standard mechanically controlled air- valve motion of the Riedler 
design is described in Chapter IX. 

Several other forms of discharge valve will be noted later, in 
connection with mechanically controlled valve motions. 

Discharge Area for Air Cylinders. The volume of air to be 
discharged from the cylinder having been reduced by compression 
to a small fraction of the volume occupied at atmospheric pressure, 
it might appear that the total area of the discharge valves could be 
made much smaller than the inlet area, without producing ex- 
cessive frictional resistance. But the compressed air must be 
forced out of the cylinder in a relatively short period of time. While 
the air enters throughout nearly the entire stroke, the delivery must 
take place while the piston is making, say, the last third or quarter 
of the stroke. Therefore, in a compressor of ordinary design, with 
several poppet inlet and discharge valves, the total discharge area 
should be about equal to the inlet area, provided the piston speed 
be moderate. When the inlet area is concentrated in a single 
valve (for example, like that of the IngersoU-Sergeant piston inlet), 
the discharge area is made about double the inlet area, though this 



I 



DISCHARGE OR DELIVERY VALVES 14! 

relation varies in cylinders of different sizes, being proportionately 
greater in the larger compressors. Obviously, other things being 
equal, the discharge area should increase with the piston speed. 
For a speed of 300 feet per minute, the best results are obtained 
by making the discharge area, say, 10 per cent, of the cylinder area; 
for speeds of 450 to 500 feet per minute, 15 per cent.* In some 
compressors, however, the discharge area is as small as from 8.5 
to 9.5 per cent. 

The above considerations apply in a measure also to the passages 
through which the air passes from the discharge valves to the pipe 
leading to the receiver. In some designs these are too restricted 
to permit a free flow of the air. The velocity of discharge should 
be made as small as possible, to minimize the resistance due to 
friction ; otherwise, during the period of delivery the pressure of 
the compressed air in the cylinder will rise momentarily above the 
normal, and then drop back after the air has passed out to the 
receiver. This causes a loss of power and unnecessary strains on 
the moving parts of the compressor. The amount of loss from 
this cause is represented by the irregular area of the a ir card which 
lies above a horizontal line drawn through the point corresponding 
to the pressure at the end of delivery (see Fig. 61). When the dis- 
charge valves first open, the piston is moving at a high velocity, 
and equilibrium with the receiver pressure is only attained as this 
velocity decreases toward the end of the stroke. 

* W. L. Saunders, " Compressed Air," Dec, 1896, p. 153. 



CHAPTER IX 

MECHANICALLY CONTROLLED VALVES AND VALVE 

MOTIONS 

The disadvantageous features of inlet valves whose opening 
and closure depend primarily upon difference of air pressure have 
led to the introduction of numerous mechanically controlled 
•valves. By their use fewer valves are required, as a rule, because 
they may be made much larger and have a higher lift. As dis- 
tinguished from ordinary poppet valves, they are operated or con- 
trolled by being in some way connected with the rotary or recipro- 
cating parts of the compressor. A prompter opening is thus 
secured, so that the compressor is enabled to take more nearly a 
full cylinder of air at each stroke. 

In some designs the connection between the valves and their 
operating mechanism is absolutely positive and fixed for any one 
setting of the valves, which are timed with respect to the piston 
stroke, so as to open at the instant the clearance air has been re- 
expanded to atmospheric pressure, and to close at the end of the 
stroke. Other designs involve the use of springs, which modify 
to some extent the operation of the controlling mechanism, thus 
allowing for variations in working conditions, as well as for in- 
accuracies of adjustment or slight derangements caused by wear 
of parts. Still other valve motions exert a partial control, which, 
within narrow limits, leaves the valve free to act under difference of 
air pressure inside and outside of the air cylinder. 

As a rule, in the recent designs of mechanical valve motions the 
inlet valves only are positively controlled, and in most cases the 
type of valve used is a modified form of the Corliss. But while 
mechanically controlled valves are often employed for the low- 

142 



I 



MECHANICALLY CONTROLLED VALVES AND VALVE MOTIONS I43 

pressure cylinders of stage compressors, they are not suitable for the 
high-pressure cylinders; the inlets of these are subjected to heavy 
pressures on both sides, and are best allowed to open and close 
solely under the difference between these pressures, which is more 
than sufficient to produce prompt action of the valve at the proper 
time. Poppet valves are therefore generally used for this service. 

Mechanical Control for Discharge Valves. The adoption of any 
system of mechanical control for discharge valves is a matter of some 
difficulty, because of the fluctuations of receiver pressure under 
which these valves are compelled to operate. In attempting to 
open them by a positive mechanical movement, at a fixed point 
of the stroke, two cases may occur: i, in event of a drop in receiver 
pressure below the normal, the valves and their controlling mech- 
anism would be subjected to a heavy strain, before the point of 
opening is reached, due to the excess of cylinder pressure; and, 
2, if the pressure in the receiver should rise above the normal, 
the valves would be held forcibly on their seats, by the excess of 
receiver pressure, after being released by the controlling gear. 
In either case, derangement or breakage of the valves or of some 
part of the controlling mechanism may occur. 

Hence, in order that the discharge valves may adjust themselves 
automatically to the varying conditions, some degree of freedom 
as to their time of opening must be allowed. It is true that the. 
range of fluctuation in receiver pressure is lessened by the use of 
air-pressure regulators (Chapter XII) ; nevertheless, only a partial 
mechanical control of these valves is practicable for any service in 
which the consumption of air is variable or intermittent. More- 
over, Corliss valves of the ordinary patterns used for compressors 
do not serve well for discharge valves where the ratio of compres- 
sion is greater than, say, three or three and one-half ; because they 
must then be set to open too late in the stroke to permit a free 
discharge. This applies to single-stage compressors, as designed 
for ordinary service, as well as to the high-pressure cylinders of two- 
stage machines. A number of devices have been introduced for 
dealing with these conditions; such as the use of relief valves 
working in conjunction with mechanically operated discharge 



144 COMPRESSED AIR PLANT 

valves; or, as in one form of the Riedler compressor, the opening 
of the valve is governed in part by the air pressure, a very small 
free lift being allowed by the controlling mechanism for affording 
the necessary relief. 

Valve Motion of Norwalk Compressor. An adaptation of the 
Corliss valve gear has been used for many years for the low-press- 
ure cylinder of this compressor (Fig. 73). One large inlet and one 
discharge valve are set in chests at each end of the cylinder. Pop- 
pet valves are employed for the high-pressure cylinder. These are 
shown in Fig. 8, together with the cross-sections of the low-pressure 
Corliss valves in their repective chests. The main valve-rod, a, 
Fig. 73, is driven by a drag- or return-crank, h, mounted on the 
crank-pin of the fly-wheel. The rod is pin-connected to a short 
lever, c, on the spindle of the forward inlet valve, and from this 
lever a link, d, passes to a corresponding connection with the inlet 
valve at the other end of the cylinder, the parts being so adjusted 
that one valve opens as the other closes. A positive movement of 
the valves is thus obtained. 

An essential feature of this valve motion is the introduction of 
the cams / / and g g, for operating the discharge valves. These 
cams are mounted in pairs on the respective inlet and discharge 
valve spindles, and form part of the short levers c. As each inlet 
valve oscillates, its cam rolls smoothly upon that of the discharge 
valve above it, the shape of each pair of cams being such that the 
discharge valve is opened full at the proper point of the stroke, i.e., 
when the pressure within the cylinder becomes equal to that in the 
discharge passage outside. Then, at the end of the stroke, when 
the cams move in the opposite direction, and while still rolling upon 
each other, the discharge valve is closed without shock by the con- 
necting link, e. This link is elastic, being made of two telescoping 
parts, somewhat on the principle of a dash-pot, thus allowing the 
freedom of movement necessary for dealing with variable receiver 
pressure. 

In recent years, a number of other compressor-builders have 
adopted modifications of the Corliss valve gear for the air cylinders. 

Nordberg Valve Motion. For single-stage compressors of this 



MECHANICALLY CONTROLLED VALVES AND VALVE MOTIONS 145 




a, 

B 
o 
U 



u 



146 



COMPRESSED AIR PLANT 



make, the inlet valves are of the Corliss type, poppets being used 
for discharge (see Fig. 46). The inlet valves are operated posi- 
tively from a triple wrist-arm, on the side of the air cylinder. This 
wrist-arm is driven by an eccentric on the crank-shaft, the con- 
necting rod being supported by an intermediate carrier arm, sus- 
pended from the engine frame. Connecting links pass from the 
wrist-arm to the valve levers. The lap of the valves can be altered, 
when necessary, by slightly shifting the angular position of the 
lever with respect to the valve spindle on which it is mounted. 
This is done by means of a pair of adjusting screws on the hub of 




Fig. 74. — Air Cylinder of Nordberg Compressor. 

the valve lever, the correctness of the valve motion being unaffected. 
The discharge valves are of the cup-poppet type. 

The double- and triple-stage compressors are provided with 
similar inlet valves, a modified Corliss valve, with double ports, 
being used for discharge (Fig. 74). This valve is shown open on 
the right- and closed on the left-hand end of cylinder. An opening 
in the center of the valve allows the air to discharge on both sides. 
In the axis of each Corliss valve are set a series of spring poppets, 
which act as relief valves when the receiver pressure falls below the 
normal. It will be seen, by reference to the above figures, that the 



i 



MECHANICALLY CONTROLLED VALVES AND VALVE MOTIONS I47 

water-jacketed area on the cylinders is unusually large, jackets 
being applied wherever possible on the heads and around the 
valves, as well as on the cylinder barrels. 

Another form of Nordberg valve gear is used in compressors in- 
tended to be driven at constant speed — by belting or gearing from 
an electric motor, water-wheel, or engine used for other service. 
While the general construction of the air cylinders is the same as 




Suction 
Fig. 75. — Section of Air Cylinder. Laidlaw-Dunn-Gordon Co. 

shown in Fig. 46, the inlet valves are provided with a releasing 
mechanism. The valves are opened and closed as usual by wrist- 
plate links, when the air pressure is normal. But when the press- 
ure increases the inlet valves are released from the valve motion 
and held open until the pressure drops; that is, the compressor is 
unloaded for the time being, useful work ceasing. The release is 
effected by introducing knock-off cams, similar to those used for 



148 



COMPRESSED AIR PLANT 



Corliss steam valves, these cams being operated by a loaded plunger 
to which the compressed air is admitted vv^hen the pressure exceeds 
the normal. With this gear the compressor is self-regulating to 
within small limits. For duplex compressors, added delicacy of 
regulation is obtained by designing the knock-off cams to unload in 
four successive steps, according to the variation in air pressure. 

Laidlaw-Dunn-Gordon Valve Motions. Several forms of me- 
chanical valve motion are made by these builders. One of them 
is shown in the arrangement of its valves, by Fig. 75, the inlet valves 




Fl3. 76. — "Cincinnati" Valve Gear, Laidlaw-Dunn-Gordon Compressor. 



being of the usual Corliss pattern, with spring poppets for the 
discharge. 

Another design, recently brought out, is the " Cincinnati" air 
valve gear (Fig. 76. See Fig. 23 also, for a plan and longitudinal 
section of the complete com.pressor). This valve motion is peculiar 
in the fact that a single Corliss valve, at each end of the cylinder, 
serves as both inlet and discharge. The cross-section of the valve, 
therefore, differs materially from the usual form, as shown by the 



MECHANICALLY CONTROLLED VALVES AND VALVE MOTIONS 149 

cut. The valve at the right-hand end of the cylinder is in position 
for admitting inlet air, the air passage being indicated by dotted 
lines; while that at the left is open for discharge, the corresponding 
inlet being closed. A large poppet is set vertically just above the 
Corliss valve. The latter is timed to open the port sufficiently early 
in the stroke to leave the poppet free to rise whenever the pressure 
in the cylinder exceeds receiver pressure. At the end of the stroke 
the Corliss valve takes its inlet position (right-hand of cut), and at 
the same time, by shutting off the discharge, confines a small 
quantity of compressed air in the passage under the poppet. This 
air acts as a cushion, and allows the poppet to seat itself slowly and 
without shock, during the return stroke of the piston. The neces- 
sity for the usual sharp closure of the discharge is thus avoided; 
the spring may be made lighter, and the wear of both valve and 
seat is reduced. 

It will be seen that the fixed mechanical control of the valves 
is exerted at three points : opening and closing of the inlet, and 
closing of the discharge. In permitting the poppet to open freely 
by the combined action of both valves, one of the chief difficulties 
of applying mechanical control to discharge valves is eliminated, 
viz. : that of dealing with the variable receiver air pressure. This 
valve motion is well suited for running at high piston speeds, as, 
for example, in the case of compressors driven by direct-connected 
motors. 

Allis-Chalmers Valve Motions. These are of several types, 
resembling in part some of those already described, but differing 
in many details. Fig. 77 shows a standard form for the duplex 
compressor, in which the Corliss inlet valves are operated from a 
triple wrist-arm, driven by an eccentric on the fly-wheel shaft. 
The discharge valves (five in number for ordinary sizes of com- 
pressor) are spring-poppets of the cup form. 

Another design of discharge valve employed by these builders 
consists of a light cup-shaped poppet, without a spring, which is 
permitted to open freely, according to the air pressures, but is 
closed positively by a plunger actuated from a separate wrist-plate 
and eccentric. A single valve is placed in each cylinder head. 



I50 



COMPRESSED AIR PLANT 



The plunger, carried by exterior guides, works within the valve and 
is so timed that it forces the valve to its seat just at the end of the 
stroke. On the return stroke of the piston the plunger recedes, 
while the valve is held on its seat by the receiver pressure until the 
pressure within the cylinder rises sufficiently to open it. In closing 
the valve, the advancing plunger is cushioned on the air in the cup 
of the valve, so that the latter is seated without shock. 

Still another form of Allis- Chalmers valve-gear consists of me- 
chanically operated Corliss valves for both inlet and discharge. 




Fig. 77. — Standard Air Valve Motion. Allis-Chalmers Co. 

The time of closing of the discharge valve is adjusted for the 
maximum working pressure. To allow for variations, small 
auxiliary spring poppets are provided, to act as relief valves. 
These open freely when the receiver pressure falls below the 
working pressure for which the positively operated Corliss valves 
are set. 

Sullivan Valve-Motions. In the cross-compound, two-stage 
compressors of this make, Corliss valves are employed for the in- 
take of both low- and high-pressure cylinders. Fig. 78 is a 
longitudinal section of the high-pressure cylinder, the discharge 



MECHANICALLY CONTROLLED VALVES AND VALVE MOTIONS 151 

valves of which are of the poppet form. Corliss discharge valves 
are used in the intake cylinder, but are accompanied by poppet 
relief valves, similar in principle to those in one of the Allis-Chal- 
mers designs described above. The air valve gear is driven by the 
usual eccentric and wrist-plate motion. 

Both air cylinders of the Sullivan two-stage, straight-line com- 
pressor are fitted with Corliss inlet valves, operated from an eccen- 
tric pin attached to the main crank-pin. The discharge valves in 




Fig. 78. — Sullivan Air Cylinder, showing Corliss Inlet Valves. 



this compressor are arranged in a rather unusual manner, being 
placed in the lower, instead of the upper, part of the cylinder 
heads. In some of the other patterns of these makers, the poppet 
discharge valves are set radially around the upper part of the cylin- 
der-head castings. (See, for example, the cross-section of two- 
stage compressor. Fig. 54.) By this arrangement the piston clear- 
ance can be made very small, and the valves, placed in removable 
seats, are surrounded by the water-jackets. 



152 



COMPRESSED AIR PLANT 



Other Mechanically Controlled Valve Motions, resembling in 
principle those already noted, though differing in details of con- 
struction, are employed in a number of compressors which need 
not be described here, such as the Franklin, Clayton, Rix, American, 
etc. With but one exception, the mechanically controlled air 
valves referred to in the preceding pages are modifications of the 
Corliss rotary or oscillating valve. A wholly different type, how- 
ever, is found in the 

Riedler Air- Valve Motion. This ingenious valve motion has 
undergone several radical modifications since it was introduced, 
about twenty years ago. Its present design, as built until recently 
by the Allis-Chalmers Co., is illustrated in Figs. 79, 80, and 81. 
Fig. 79 comprises side and half -end elevations of the low-pressure 
air cylinder. The mechanical control is exerted through a wrist- 
plate. A, supported on the side of the cylinder and operated through 
a lever from an eccentric on the fly-wheel shaft. Back of the wrist- 
plate is a horizontal sliding plate, B, to which the links, E, E, are 
pinned. This plate is caused to reciprocate, through a distance 
equal to the permitted lift of the valves, by two cams cut on the 
periphery of the wrist-plate and working against studs set in B. 
The motion thus transmitted through the links, E, oscillates the 
transverse rock-shafts, D, and produces the necessary throw of the 
forked levers, C (Fig. 80), which control the closure of the valves. 
The rock-shafts are carried in bearings outside of the cylinder- 
head housings. 

The four valves, two suction and two delivery, are almost iden- 
tical in design, consisting of an annular seating portion, connected 
by radial ribs to the central disks. They are of a double-seated 
poppet type, the air passing within the seating ring as well as 
around its periphery. Screwed into the valve is a long stem, 
passing out through a stuffing-box in the cylinder head and into a 
bonnet bolted on outside. Within the bonnet is a dash-pot whose 
piston is attached to the valve stem. 

The operation of the inlet valve, F, Fig. 80, is as follows: At 
the beginning of the stroke the forked lever, C, is depressed by the 
rock-shaft and link, noted above. This leaves the valve free to 



MECHANICALLY CONTROLLED VALVES AND VALVE MOTIONS 1 53 




154 



COMPRESSED AIR PLANT 





MECHANICALLY CONTROLLED VALVES AND VALVE MOTIONS 1 55 





156 



COMPRESSED AIR PLANT 



open, its movement being steadied by the dash-pot piston, G. 
The resistance presented by this piston is regulated by the adjusting 
screw, H. For ordinary sizes of compressor the total lift of the 
valve is one inch, giving a large area of opening. (The icj-in. 
valve shown in the cut, which is for the low-pressure cylinder of a 
2/^' and 38'' X A^" compressor, has an area of 45 sq. ins.) 
Toward the end of the stroke the forked lever begins to rise, there- 
by bringing the valve gradually nearer its seat, as the piston velocity 
decreases. In completing its movement the lever forces the valve 
upon its seat promptly at the end of the stroke. By this device, the 
valve attains its maximum lift and area of opening toward the 
middle of the stroke, when the velocity of the inflowing air is great- 
est, and is brought nearer its seat as the flow diminishes, so that the 
complete closure is effected instantaneously at the proper time. 

A similar control is exerted over the delivery valve, though the 
details of its bonnet, dash-pot, and forked lever are quite different, 
as shown by Fig. 81.* At the proper point of the stroke the 

lever is depressed, so that the 
valve is free to open when the 
air pressure in front of the 
advancing piston has reached 
receiver pressure. Then, as 
the velocity of outflow dimin- 
ishes toward the end of the 
stroke, the valve is forced 
nearer its seat, and a prompt 
closure takes place the instant 
the stroke reverses. As a re- 
sult of this mechanical control, together with the action of the 
dash-pot, the operation of the Riedler valves is attended with but 
little shock, thus permitting a high piston speed. 

Cam-Controlled Inlet Valve. At the Lens colliery, in France, 
a cam movement has been successfully applied for controlling the 
opening of a poppet inlet valve (Fig. 82). The stem of the valve 

* The delivery valve in the cut is the same size as the suction valve previously 
described. It is designed for a smaller compressor, 15" and 24" X 36". 




Fig. S2. — Cam-Controlled Inlet Valve. 



MECHANICALLY CONTROLLED VALVES AND VALVE MOTIONS 1 57 



is provided with a spiral spring, and projects from the cylinder 
head, as usual. At the beginning of the stroke the valve is opened 
rapidly by a cam, a, of peculiar shape, playing against the end of the 
stem. The cam is mounted upon a small shaft which is geared to 
revolve once for each revolution of the compressor. At the end of 
the stroke the cam allows the valve to close under the action of the 
spring.* 

Sturgeon Inlet Valve. This peculiar valve, of an air com- 
pressor made in England, furnishes an example of a positive 
movement entirely different in principle from those already de- 
scribed. It is a large annular valve, c (Fig. S;^), encircling the piston 
rod in each cylinder head, and 
is operated directly by the 
movements of the rod it- 
self. | The connection be- 
tween the valve and rod is 
frictional only, being brought 
about by a gland, e, which 
serves also to form a stufhng- 
box for the piston. By tight- 
ening or loosening the nuts, 
a, of the bolts by which the 
gland is attached to the valve 
flange, any desired amount of 
grip upon the piston rod can 
be obtained. This frictional 
grip is regulated so that the valve will not be opened until the 
clearance air has been re-expanded nearly to atmospheric pressure. 
Flanges on the ends of the valve limit its play in each direction, 
controlling the amount of lift and area of opening. The valve 
and stuffing-box together form the bearing of the piston rod in 
each cylinder head. At the end of the stroke a recess in the 
piston receives the large inner flange of the valve so as to 
diminish the clearance. The mechanism is simple and its work 
satisfactory. 

* H. W. Hughes, "Text-book of Coal Mining," p. 55. t Idem, p. 53. 




Fig. 83. — Sturgeon Inlet Valve. 



158 COMPRESSED AIR PLANT 

An arrangement resembling the above has been used in a 
compressor made by the Dover Iron Co., Dover, N. J. 

The well-designed Koster valve is of the piston type, almost 
unknown in this country for air-compressor service. It is now em- 
ployed by several European makers, among them: Pokorny & 
Wittekind, Frankfort, Neumann & Esser, Aachen, and W. H. 
Bailey & Co., Manchester, England. The valves, both inlet and 
discharge, are very large in area and are mounted on a longitudinal 
spindle deriving its reciprocating motion from an eccentric on the 
crank-shaft. Positive opening and closure are imparted to the 
inlet valves, but the opening of the delivery port is effected by an 
independent poppet, encircling the spindle and provided with a 
light spring. This valve motion constitutes a highly developed 
type, and is both reliable and efficient. 



CHAPTER X 

PERFORMANCE OF AIR COMPRESSORS* 

The performance or duty of air compressors may be designated 
in several different ways. 

First. A standard of rating, useful for ordinary purposes, is 
the duty in terms of cubic feet of free air compressed per minute 
to a given pressure; or, the volumetric output. The theoretical 
output is found by multiplying the net piston area in square feet 
by the distance travelled by the piston in feet per minute. The 
actual output will be less than the theoretical on account of vari- 
ous losses due to leaks, clearance, induction of warm air, friction 
of inlet valves, etc. In a properly designed compressor an allow- 
ance of fifteen per cent, to eighteen per cent, is sufficient to cover 
these losses, which must not be confounded with the mechanical 
loss of work — that is, the work expended in overcoming the fric- 
tion of the compressor — and the loss of useful work due to the 
heating of the air under compression. 

Having found the capacity of the compressor, in terms of cubic 
feet of free air, the volume, V, occupied by this air at any given 

VP 

pressure, P', is calculated by the formula already given: V = -^, 

in which the following values are now assigned, viz. : 

V = initial volume of given quantity of air in cubic feet. 

P = normal absolute pressure of atmosphere (14.7 lbs.). 

P' = absolute pressure of air under compression, i.e., gauge 
pressure -f- 14.7 lbs. 

For example, 100 cu. ft. of free air, compressed isothermally to 
65 lbs. gauge pressure, will occupy a volume: 

_-, 100 X 14-7 o r^ 

V = -'^ = 18.4s cu. ft. 

65 + 14.7 

* The deductions of the work formulae used in this chapter are given in detail 
in Chapter III. 

I. "9 



l6o COMPRESSED AIR PLANT 

Conversely, the volume of free air represented by 18.45 c^- f^- 
of air at 65 lbs. gauge pressure is : 

^, YV 18.45 (65 + 14.7) u 

V = ^ = — ^ ^ =^^ = 100 cu. ft. 

P 14.7 

By applying the 15 to 18 percent, allowance for losses stated 
above, this allowance depending on the type of compressor, re- 
sults are obtained sufficiently accurate for practical purposes. 
As the volumetric output of a compressor of given size of cylinder 
depends on the density of the intake air, it will obviously be 
reduced when working at an altitude above sea-level. (See Chap- 
ter XIII.) 

Second. The size of the compressor may be designated in 
terms of the horse-power developed by the steam end, indicator 
cards being taken while running at normal working speed and while 
the usual volume of air is being consumed. 

Third. The effective horse-power of the quantity of com- 
pressed air delivered is determined from an indicator card, taken 
from the air cylinder. In testing a compressor it is customary 
to take a series of cards, simultaneously from both ends of the 
steam and air cylinders. They may then be compared, as shown 
by Fig. 24. 

If indicator cards be not available, the theoretical horse-power 
for single-stage adiabatic compression may be calculated by the 
formula : 

33, ooo(w — i) L \ P / J 

P = normal atmospheric pressure per square inch (14.7 pounds). 

P' = final absolute pressure per square inch. 

V = the volume of free air compressed per minute, in cubic feet. 

n = exponent of the compression curve, as given under the the- 
ory of air compression, viz.: for adiabatic compression, w = 1.406, 
and varies down to x.i8 or 1.2, depending on the efficiency of the 
cooling arrangements. For the best single-stage compressors, n = 
say, 1.25 or 1.3. 

For isothermal compression , the expression for the horse- power is : 



PERFORMANCE OF AIR COMPRESSORS 



l6l 



H.P. = 



144 



P v(Nap. log. ^) 



33,000 

Table V shows the horse-powers required, under the conditions 
named, to compress one cubic foot of free air per minute to different 
gauge pressures, by single-stage compression : 



Table V 







Single-Stage Compression, from Atmospheric 






Pressure at Sea-Level. 


Initial Temp., 60° Fah. 






Horse-Power Required 


to Compress i cu. ft. of 






Free 


Air. 




Atmos- 








pheres Ab- 
solute, or 
Ratio of 






Gauge 

Pressure, 

Lbs. 


Theoretical Horse-Power. 


Actual Horse- Power (Approx.). 


Compres- 








sion. 






Allowance for 


Allowance for 






Isothermal 


Adiabatic 


Losses above 


Losses above 






Compression. 


Compression. 


Adiabatic Com- 
pression, is%. 


Adiabatic Com- 
pression, 20^. 


20 


2.36 


-0551 


.0626 


.0720 


-0751 


25 


2.71 


.0637 




0741 


.0852 


.0890 


30 


3-04 


■07^3 




0843 


.0970 


.1011 


35 


3-3^ 


.0782 




0941 


.1082 


.1129 


40 


3-72 


.0842 




1029 


.1183 


.1234 


45 


4.06 


.0895 




1115 


.1282 


-^33^ 


50 


4.40 


.0950 




1191 


.1370 


.1430 


55 


4-74 


.0994 




1269 


-1460 


.1522 


60 


5.08 


.1041 




1337 


-1537 


.1604 


65 


5-42 


.1081 




1401 


.1610 


.1681 


70 


5-76 


.1123 




1468 


.1690 


.1761 


75 


6.10 


.1162 




1535 


•1765 


.1842 


80 


6.44 


-I195 




1591 


.1830 


.1910 


85 


6.78 


.1224 




1 65 1 


.1900 


.1961 


90 


7.12 


.1256 




1703 


•1955 


.2040 


95 


7-46 


-1287 




1760 


.2024 


.2112 


100 


7.80 


-1315 




1807 


.2080 


.2168 


no 


8.48 


.1366 




1894 


.2180 


.2272 


125 


9-50 


.1442 




2025 


.2328 


.2430 



In columns three and four of Table V are shown the theoretical 
horse-powers required for isothermal and adiabatic compression. 
The results of isothermal compression are wholly unattainable in 
practice, and are placed here only for purposes of comparison. 
They represent an ideal which it is desirable always to keep in view. 

* The Naperian or hyperbolic logarithm of a number is obtained by multiplying 
the common logarithm by the constant 2.302585. 

11 






1 62 COMPRESSED AIR PLANT 

The figures given in the column of adiabatic compression are based 
on the assumptions that there is no radiation of heat from the air 
cylinder, and that the temperature of the air after delivery has 
become normal, its volum.e being therefore reduced to that which 
is practically available for use. No allow^ances are included in 
these figures to cover losses other than that due to the heating of the 
air under compression. But the full amount of loss represented by 
adiabatic compression can never be suffered in the operation of com- 
pressors, however imperfect their design. The actual compression 
line must always be lower than the adiabatic line, because of the 
radiation of heat through the cylinder walls. In ordinary, single- 
stage compressors, properly water- jacketed and run at a reasonable 
piston speed, the compression line falls considerably belo^ the 
adiabatic line. Whatever diminution of loss is effected by cooling 
of the air in the cylinder may therefore be credited against the other 
unavoidable losses, partially offsetting them, viz.: frictional or 
mechanical loss in the compressor, friction of inlet valves, heating 
of the intake air by contact with the hot metal surfaces, and piston 
clearance of the cylinder. These losses are variable in amount, 
depending on the design of the compressor. 

In the absence of indicator cards, giving the actual results in 
individual cases, estimates based on practice may be made of the net 
power loss experienced in operating compressors, which will be con- 
venient for reference. With this understanding, an attempt is 
made, in columns five and six of the above table, to show the actual 
horse-power required to compress one cubic foot of free air, under 
the conditions stated at top of the columns. Thus, in column five^ 
fifteen per cent, is assumed as a fair estimate, in case of well- 
designed and operated single-stage compressors, of the additional 
power required, over and above that for theoretical adiabatic com- 
pression ; this fifteen per cent, being taken as : the loss in purely 
adiabatic compression, minus the effect of ordinary water-jacket 
cooling, plus the other four losses mentioned at end of preceding 
paragraph. In column six, the power consumed in adiabatic 
compression is increased by twenty per cent., which represents 
relatively poorer work. 



PERFORMANCE OF AIR COMPRESSORS 163 

The figures in columns 3 and 4 or 5 and 6 (which are for 
free air), if multiplied by the corresponding ratios of compression 
(column 2), will give the respective theoretical and actual power 
costs of furnishing one cubic foot of compressed air, at the gauge 
pressures stated. 

Work Done by Stage Compressors. The theoretical horse- 
power required to compress a given volume of free air to any given 
pressure, P', is computed for a two-stage compressor by the for- 
mula : 



^^,000 w— iL^P/ -J 



H.-P. = 

33,000 

This formula is derived from that for single-stage compression 
by dividing equally between the cylinders the total work done, and 
then taking the sum of the two. 

For three-stage compression the formula becomes : 

33,000 n—i L \ P / J 

Reducing the constants, and for a volume of one cubic foot of 
free air, these formulas may be simplified thus : 



Two-stage, H.-P. = 0.449 [_ y-p j -" i J 

I — / p^ \ 0.062 —1 
Three-stage, H.-P. = 0.6735 [(|-) -ij 



In these expressions it is assumed that the work of compres- 
sion in each cylinder is done adiabatically, and that the temper- 
ature of the air after leaving the cylinder is reduced by intercool- 
ing to the initial temperature. 

For convenience, the horse-powers for stage compression at sea- 
level, both theoretical and actual, are given for a few gauge press- 
ures in the following table; the figures in the fifth and seventh 
columns being taken as an approximation to the results obtainable 
in practice from stage compressors of the usual designs. 

At elevations above sea-level, P is less than 14.7, and for any 
given altitude the atmospheric pressure must therefore be known. 



164 



COMPRESSED AIR PLANT 

Table VI 





Ratio of Com- 


HORSE- 


Power per Cubic Foot of Free Air. 


Gauge 




Two-Stag^e 
Compression. 


Three-Stage 
Compression. 


Pressure, 


pression 
= P^ 








Lbs. 


Isothermal 
Compression. 


Adiabatic 


Actual H - 
P., on basis 


Adiabatic 


Actual H.- 
P., on basis 








Compres- 


of Adia. 


Compres- 


of Adia. 








sion. 


Comp'n 

-t- 1%%. 


sion. 


Comp'n 












+ 15^. 


70 


5-76 


0.1123 


0.129 


0.152 






80 


6-44 


-1 195 


.138 


-163 






90 


7-12 


.1256 


.147 


■ -^n 






100 


7.80 


-1315 


-154 


.182 


0.145 


0.167 


120 


9.16 


.1420 


.169 


.199 


-158 


.182 


140 


10.^0 


.1508 


.181 


.213 


.169 


.194 


160 


11.88 


-1583 


.192 


.226 


.179 


.206 


180 


13-24 


-1654 


.202 


.238 


.188 


.216 


200 


14.60 


.1720 


.212 


.250 


.196 


.225 


250 


18.00 


-1853 


.232 


.274 


.213 


-245 


300 


21.40 


.1963 


-249 


.294 


.228 


.262 


350 


24.80 


.2058 


.264 


.312 


.241 


•277 


400 


28.20 


.2140 


.277 


•327 


.252 


.290 


450 


31.62 


.2215 


.289 


•341 


.262 


.301 


500 


35-01 


.2280 






.271 


-311 


550 


38.41 


-2339 






.280 


.222 


.- 600 


41.80 


-2393 






.288 


•331 


650 


45.21 


-2443 






-295 


-339 


700 


48.62 


.2490 






.301 


-346 


800 


55-42 


-2574 






-314 


.361 



Table VII will be found useful for making calculations in 
which are used volumes and mean cylinder pressures for isother- 
mal and adiabatic compression. 

In this table the mean pressures per stroke, given in the 
fifth and sixth columns, are obtained from the formulas for iso- 
thermal and adiabatic single-stage compression, which precede 
Table V, except that they are here expressed in terms of foot- 
pounds of work, instead of horse-power. These formulas may be 
put respectively in the following forms : 

Mean pressure per stroke (isothermal) = P X Nap. log. — 

Mean pressure per stroke (adiabatic) =3.463 P I ( ~p ) — i J 



n 
* The constant 3.463 = = 



1.406 

.406 



PERFORMANCE OF AIR COMPRESSORS 



i6s 



Table VII* 



1 

1 




0) 

^ c 

ScoJ 
sob 




'0 


:ean Pressure per 
Stroke; Air at Con- 
stant Temperature. 
Pounds. 


m 


ill 

III 


O 


< > 


> 


> 


S 


S 


H 





I I 




I 








60° 


I 


1.068 


9363 


.9500 


.96 


-975 


71 


2 


1. 136 


8803 


.9100 


1.87 


1. 91 


80.4 


3 


1 . 204 


8305 


.8760 


2.72 


2.80 


88.9 


4 


1.272 


7861 


.8400 


3-53 


3-67 


98 


5 


1.340 


7462 


.8100 


4-30 


4-50 


106 


lO 


1.680 


5952 


.6900 


7.62 


8.27 


H5 


15 


2.020 


4950 


.6060 


10.33 


II. 51 


178 


20 


2.360 


4237 


-5430 


12.62 


14.40 


207 


25 


2.700 


3703 


.4940 


14-59 


17.01 


234 


30 


3.040 


3289 


-4538 


16.34 


19.40 


252 


35 


3-381 


2957 


.4200 


17.92 


21.60 


281 


40 


3-721 


2687 


-3930 


19.32 


23.66 


302 


45 


4.061 


2462 


.3700 


20.57 


25-59 


321 


50 


4.401 


2272 


-3500 


21.69 


27-39 


339 


55 


4.741 


2109 


-3310 


22.76 


29.11 


357 


60 


5.081 


1968 


.3144 


23-78 


30-75 


375 


65 


5-423 


1844 


.3010 


24.75 


32-32 


389 


70 


5.762 


1735 


.2880 


25.67 


33-83 


405 


75 


6.102 


1639 


.2760 


26.55 


35-27 


420 


80 


6.442 


1552 


.2670 


27.38 


36-64 


432 


85 


6.782 


1474 


.2566 


28.16 


37-94 


447 


90 


7.122 


1404 


.2480 


28.89 


39-18 


459 


95 


7.462 


1340 


.2400 


29-57 


40.40 


472 


100 


7.802 


1281 


.2320 


30.21 


41.60 


485 


105 


8.142 


1228 


.2254 


30.81 


42-78 


496 


no 


8.483 


1178 


.2189 


31-39 


43-91 


507 


115 


8.823 


^^33 


.2129 


31.98 


44-98 


518 


120 


9.163 


1091 


.2073 


32-54 


46.04 


529 


125 


9-503 


1052 


.2020 


33-07 


47.06 


540 


130 


9-843 


1015 


.1969 


33-57 


48.10 


550 


135 


10.183 


0981 


.1922 


34-05 


49.10 


560 


140 


10.523 


0950 


.1878 


34-57 


50.02 


570 


145 


10.864 


.0921 


-1837 


35-09 


5 1 . 00 


580 


150 


11.204 


.0892 


.1796 


35-48 


51.89 


589 


160 


11.880 


.0841 


.1722 


36.29 


53-65 


607 


170 


12.560 


.0796 


-1657 


37-20 


55-39 


624 


180 


13.240 


-0755 


.1595 


37-96 


57-01 


640 


190 


13.920 


.0718 


.1540 


38.68 


58-57 


657 


200 


14.600 


.0685 


.1490 


39-42 


60.14 


672 



* Kents' " Mechanical Engineers' Pocket Book, 
ards' " Compressed Air," p. 20. 



Taken from a table in Rich- 



1 66 COMPRESSED AIR PLANT 

The work done during one stroke of the compressor is found 
by multiplying the mean pressure by the volume in cubic feet, 
V, traversed by the piston. 

When air is compressed adiabatically, the relation between the 
temperature T, of the air at the beginning of compression, and the 
temperature at the end, T', is shown by: 

T=M . whence T' = tQ 

The final temperature may also be found from the formula: 



T 



■-^(ip 



T and T^ being absolute temperatures and P, P' absolute pressures 
in each case. 

The compression curve of an air-indicator card maybe con- 
structed as follows, PV being the pressure and volume at one point 
of the curve and P' V the pressure and volume corresponding to any 
other point. Designating the index number of the curve by x: 
P /V'\^ 

p7 = Vy- / * "^^^"^ ^^^^' 

log. (p^) = ^ log. (-^ ) ; whence, x = -^^ 

log. y-^) 

In considering an air card, it should be observed that the sev- 
eral lines have significations entirely different from those of a steam 
card. Referring to Fig. 84, which represents an ideal card: 
A B is the admission line, B C the compression line, C D the de- 
livery or discharge line, and D A the re-expansion line. The last- 
named line represents the effect of the re-expansion of the air filling 
the clearance space in the cylinder, on beginning a stroke (see 
Chapter V). Comparing the lines of the air and steam cards, 
they are found to be reversed, thus: 

Air Card. Steam Card. 

Admission line. Back -pressure or exhaust line. 

Compression line. Expansion line. 

Delivery line. Admission line. 

Re-expansion line. Compression line. 



PERFORMANCE OF AIR COMPRESSORS 167 

The elements of an air-indicator card, together with the work 
done, as represented by the several lines and areas, will be further 
elucidated by referring to Fig. 85 » 

In this analysis the compression is supposed to be done adia- 
batically. 

Delivery Line 



B Admission Line ^ 

Fig. 84. 

Let A D = normal atmospheric line at sea-level. 

A G=P= corresponding atmospheric pressure, acting behind 
the piston at the beginning of the stroke (neglecting 
valve resistances and effect of clearance of previous 
stroke). 
G E = A D = length of stroke of piston. 
A B = adiabatic compression curve. 
B C = delivery line. 
At the point B the useful work of compression ceases ; during 
the remainder of the stroke the volume of compressed airV^ at 
the absolute pressure?', is being forced out of the cylinder through 
the delivery valves. 

The area A B F G = the absolute work of compression. 
The area B C E F = the absolute work of delivery. 
The sum of these areas represents the total absolute work 
(that is, on the basis of absolute pressure) done during compres- 
sion and delivery. 



i68 



doMPRESSED AIR PLANT 



Area ADEG-work done for the entire stroke by atmos- 
pheric pressure behind the piston. 

Area A B H -= net work of compression. 

Area B C D H = net work of delivery. 

Area A B C D A = total net work for entire stroke. 

From this analysis another method may be derived for calcu- 
lating the theoretical horse-power required for compressing air. 




T 



-_i. 



It will often be found useful, when a table of temperatures of com- 
pression is available. 
Let w= weight of a unit volume (i cu. ft.) of free air = .0765 lbs. 

C^ = specific heat of air at constant pressure = .2375. 

Cj,= specific heat of air at constant volume = .1689. 

= 3.46. 



Q 

Whence, r^=w= 1.406, and 



J = Joule's heat unit, taken as 778 ft. lbs. The work rep- 
resented by the area A B H = 

J XwXCJT' -T) -F {V - VO. 



PERFORMANCE OF AIR COMPRESSORS 1 69 

Also, the work done during delivery = B C D H == V (P' — P). 
Hence, the total net work for one stroke of the piston 
= area A B C D A = J X 2^; X C, (T' - T) - (P V - P' VQ. 
If Cp be substituted for C^, then P V = P' V, according to the 
general equation for air compression and the total work, W = 
]XwXC^ (T' - T). 
Substituting for J, w, and C^, their constant numerical values: 

W = 14.13 IT' - T), 
or, to compress i cu. ft. of air per min., at 60° F., and at sea-level, 



H.-P. = 0.225 [^- - i] 



By referring to the last column of Table VII and remembering 
that T and T' are absolute temperatures, i.e., thermometric tem- 
peratures plus 459° F., the horse-power required for compressing 
one cubic foot of free air adiabatically to any gauge pressure may 
readily be calculated. 

Other expressions, for the mean effective pressures, may also 
be deduced from what precedes: 
M.E.P. for the entire stroke = 

■ f^.(?-)=3*K?-)-3*r[(£)=.'-.] 

M.E.P. during delivery = — (P' - P). 

The M.E.P. for compression only is found by taking the dif- 
ference between the pressures calculated by the last two formulas. 

The results obtained from the above expressions for work and 
mean effective pressure are theoretical. To find the actual horse- 
power required, allowances must be made for the several losses 
experienced in the operation of the compressor, as already set 
forth. 

Compresssor Tests. To indicate the observations required to 
secure the data for the complete test of a compressor, together 
with the deductions from the observed data, the following record 
of the test of a compound, two-stage Nordberg compressor, at 






iyo COMPJIESSED AIR PLANT 

the mines of the Tennessee Copper Co., will be found useful.* 
It will be noted that items 28, 29, and 32 to 35, are necessary in this 
case, because the boiler plant supplied steam for the hoisting en- 
gine and an independent condenser, as well as for the compressor. 
Though the hoist was not running, steam was passing continuously 
to the jackets of the cylinders. The same conditions would often 
be met in other tests. The boiler feed water was taken from a 
wooden tank, and during the run this water was supplied from 
two barrels on scales set temporarily over the tank. The water of 
condensation from steam jackets and reheater was drawn off 
continuously and also weighed. The calorimeter tests were made 
with a Peabody throttling calorimeter. Eight sets of indicator 
cards were taken during the 8-hour test, at hourly intervals. 

Items of Compressor Test 

Altitude, 1,800 feet. 

1. Date of test, February i6, 1902 

2. Duration of test, hours 8 

3. Diameter of high-pressure steam cylinder (steam jacketed), 

inches 14 

4. Diameter of low-pressure steam cylinder (steam jacketed) 

inches 28 

5. Diameter of low pressure air cylinder, inches 24^ 

6. Diameter of high pressure air cylinder, inches 15I 

7. Stroke of all pistons, inches 42 

8. Diameter of piston rods, inches 2}^^ 

9. Revolutions of engine, average per minute 90 

ID. Piston speed per minute, feet 630 

11. Steam-gauge pressure, average, pounds 145-9 

12. Temperature of steam in steam pipe, average degrees Fah 364 

13. Steam pressure in reheating receiver, average pounds 8 

14. Vacuum in condenser, average inches 25.66 

15. Air pressure in intercooler, average, pounds 22.63 

16. Air pressure in receiver, average, pounds 79.3 

17. Temperature of air at intake, average degrees Fah 65.0 

18. Temperature of air leaving low-pressure cylinder, average de- 

grees Fah 211. 5 

19. Temperature of air leaving intercooler, average, degrees Fah. . . 78.5 

20. Temperature of air leaving high-pressure cylinder, average, 

degrees Fah 240.0 

* Abstracted from an article by J. Parke Channing, Mines and Minerals, May, 
1905, P- 475- 



PERFORMANCE OF AIR COMPRESSORS 17I 

21. Indicated horse-power in high-pressure steam cylinder, aver- 

age 140.12 

22. Indicated horse-power in low-pressure steam cy Under, average. 153-03 

23. Indicated horse-power in both steam cylinders, average 293.15 

24. Indicated horse-power in low-pressure air cylinder, average. . . . 143-79 

25. Indicated horse-power in high-pressure air cylinder, average. . . 135.02 

26. Indicated horse-power in both air cylinders, average 278.81 

27. Feed-water weighed to boilers, pounds 43)343 

28. Re-heater and jacket water from compressor, weighed, pounds . 4,081 

29. Average temperature of re-heater and jacket water, degrees Fah. 356.7 

30. Total heat in i pound of steam at 356.7 degrees Fah., heat units 1,190.7 

31. Total heat in i pound of water at 356.7 degrees Fah., heat units 328.9 

32. Equivalent credit for re-heater and jacket water, pounds 1,127.00 

33. Water weighed from condensation in hoisting-engine jacket, 

pounds 1,781.00 

34. Steam used to run condenser, pounds 4,320.00 

35. Total credits to feedwater, pounds 7,228.00 

36. Total feedwater charged to engine, pounds 36,115.00 

37. Moisture in steam shown by Peabody calorimeter, per cent 1.30 

38. Credit for moisture in steam, pounds 473-oo 

39. Total steam charged to engine, pounds 35,642.00 

40. Dry steam per hour charged to engine, pounds 4,455.00 

41. Steam consumption per indicated horse-power per hour, pounds 15-19 

42. Guaranteed steam consumption per indicated horse-power per 

hour, at 92 revolutions per minute, pounds 14.00 

43. Excess of steam consumption per indicated horse-power per 

hour over guarantee, pounds 1.19 

44. Theoretical delivery of free air per minute at 90 revolutions, . 

cubic feet 2,037.8 

45. Slip of air (percentage) 3.0 

46. Actual sUp of air per minute, cubic feet 61. i 

47. Actual delivery of free air per minute, average cubic feet 1,976.7 

48. Theoretical horse-power required to compress and deliver actual 

deHvery of air at receiver pressure by adiabatic compression 306.53 

49. Theoretical horse-power required to compress and deliver actual 

delivery of air at receiver pressure by isothermal compression 229.00 

50. Actual horse-power shown by air indicator cards 278.81 

51. Actual horse-power shown by steam indicator cards 293.15 

52. Actual horse-power consumed by friction of engine 14-34 

53. Efficiency ratio between steam and air cylinders, per cent 95.1 

54. Efficiency ratio between steam and air cylinders guaranteed by 

builder, per cent 87.0 

55. Efficiency of steam, or ratio of steam indicated horse-power to 

theoretical air indicated horse-power, isothermal compres- 
sion, per cent 7^-'^ 

One of the combined indicator cards, from which the averages 
in items 21 to 26 were calculated, is shown in Fig. 86. 



COMPRESSED AIR PLANT 
Discharge Press\ire 




Fig. 86. — Combined Cards. Two-Stage Nordberg Compressor. 



' ^^ V' — 

/\ Crank End 
/ \ 



'\/\/\/\A/^-'^^^i 



Crank End ' 



L.P. Air Cylinder 



::;^ 



H.P. Air Cylinder 




Fig. 87. — Combined Cards, " Imperial Type 10," Two-Stage, Direct-Connected, 
Electrically Driven Compressor. Air Cylinders 23" and 14" X 20". 



Revolutions per minute 187 

Piston speed, feet per minute .. 623 

Discharge air pressure, lbs 93 

Intercooler pressure, lbs 24 

Volumetric efficiency (from card) 95 
l.H.P. of low-pressure cylinder . 132 



3% 



l.H.P. of high-pressure cylinder .. 120.8 

Total I. H.P 252.8 

Free air delivered per minute 

cubic feet (from card) 1706 . 

Efficiency compared with adiabatic 97-2% 
Efficiency compared with iso- 
thermal 84% 



PERFORMANCE OF AIR COMPRESSORS 



173 



In further illustration of the performance of air compressors, 
the combined card from an Ingersoll-Rand "Imperial" two-stage 
compressor, taken at one of the Berwind-White Coal and Coke 
Company's mines, is given in Fig. 87. 



-93.5"/- 
-100 i- 



h30 
25 
20 
H5 
10 
5 




YiG, 88.— Card from 3oi"X 24" L. P. Air Cylinder of Style "O," Ingersoll-Rand 
Compressor. (St. press., 115 lbs.; air press., 28 lbs.; r. p. m., 100; spring, 20.) 

Figs. 88 and 89 show shop-test cards, with accompanying 
data, from the high- and low-pressure air cylinders of an Ingersoll- 
Rand style "O" compressor. 

A Record of Field Tests. It w^ould undoubtedly tend to the 
securing of greater economy in the production of compressed 



A 



^■===-^v- 



-100 

-90 

-80 

-70 

-60 

-50 

[-40 

30 

20 

10 





Fig. 89.— Card from i8|" X 24" H. P. Air Cylinder, Style "O," Ingersoll-Rand 
Compressor. (St. press., 115 lbs.; air press., 100 lbs.; r. p. m., lop; spring, 60.) 



air, if superintendents and master mechanics would give more 
attention to the actual results produced by the operation of com- 
pressors in their charge, and study carefully the frequently un- 
favorable conditions under which these machines are called upon 
to work. 



174 COMPRESSED AIR PLANT 

Few records of the actual effective horse-power of air com- 
pressors have been published. To express the efficiency, it is 
customary to divide the horse-power of the air cylinder by the 
horse-power of the steam cylinder, as determined by indicator 
cards. The manufacturer of air compressors usually rates his 
machine on the basis of its mechanical efficiency, without taking 
into consideration any losses except those of friction. Such a cri- 
terion does not properly measure the relative commercial values 
of compressors, nor does it present any indication as to the effect- 
ive horse-power developed under ordinary working conditions. 

A series of tests were made in 1909 by Richard L. Webb, 
consulting engineer, of Buffalo, N. Y., on a large number of com- 
pressors in a well-known Canadian mining district. In conducting 
these tests, Mr. Webb had access to plants which have been in 
operation for a year or more under normal working conditions, 
and I believe his results will be of value not only to users of air 
compressors, but also to the manufacturers. As a rule, the plants 
tested were in the care of competent machinists and in good 
running order, so that the results obtained may be taken as repre- 
senting a fair average of current practice in the United States and 
Canada. The results of a few of these tests are given here to show 
the importance of determining the actual efficiency of air com- 
pressors when working under the conditions prevailing in most 
mines. 

Mode of Conducting the Tests. The following plan was em- 
ployed in each case. First, a boiler test was run for not less than 
two weeks, the coal being carefully weighed, the boiler feed water 
measured, and the total revolutions of the compressor recorded by 
a revolution counter. From these data, the cost per boiler horse- 
power and the average speed of the compressor were determined. 
Second, the compressor was operated at various speeds over its 
entire range. By means of a meter installed in the steam pipe 
near the throttle, the total steam consumed, in pounds per hour, 
was measured. Indicator cards were taken on all cylinders, together 
with temperatures at the air inlet, intercooler, and discharge. To 
measure the actual volume of air delivered, a meter was placed in 



PERFORMANCE OF AIR COMPRESSORS 



175 



the discharge pipe outside of the receiver. A number of simulta- 
neous readings on all instruments were taken at each speed. From 
these were calculated the total horse-power of the steam and air 
cylinders, the steam consumption, and the total piston displace- 
ment per minute. 

The air and steam meters were of the Dodge type, as modified 
by the General Electric Company, and were operated by their 







At\f\ 


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/ 




z 




z 




z 


9 "in 


7>^ 




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yW 




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1900 
1800 
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1600 
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1300 5-1 
1200 ^ 
1100 '< 
1000 §> 



40 a 





kg^30 O 



20 o 



10 20 30 



50 ( 
R. P. M. 



70 80 



100 110 



700 ' 
600 
500 
400 
300 
200 
100 




Fig. 90. — Compressor Plant No. i. 



expert sent for this purpose. The indicators were of the Roberts- 
Thompson and the American-Thompson make, which are well 
known and generally accepted as standard. Their springs were 
calibrated by a standard gauge. 

Results of the Tests. As was to be expected, the friction loss 
was found to be only a small item in the total. The other losses, 
which are frequently overlooked or disregarded, played a large 
part in cutting down the efficiency. The capacity of air com- 



176 



COMPRESSED AIR PLANT 



pressors is usually rated according to the volume of the cylinders. 
On this basis, the mechanical efficiency only is given. For ex- 
ample, if the horse-power of the air cylinder is 100 horse-power 
and the horse-power of the steam cylinder no, the efficiency of 











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A = Cost of Coal per I. H. P. per Annum 
B= " " , " and Labor 
C= " " " " " and Int. and Dep. 
D= Total Cost per Annum 


















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R. P. M. 

Fig. 91. — Compressor Plant No. i. 

the compressor is rated as 91 per cent. This rating disregards the 
losses due to adiabatic compression, heating of the cylinder and 
friction of the inlet and delivery valves. The tests show the 
friction loss of the engine itself to be usually not less than 10 



PERFORMANCE OF AIR COMPRESSORS 1 77 

per cent., and often considerably larger. Losses from the other 
causes mentioned were found to range from 30 per cent. up. 

As Mr. Webb is not at liberty to disclose the identity of the 
particular plants at which the tests were made, each test has 
been designated by a number. 

Test of Plant Number One. This consists of three 125 H.P. 
return tubular boilers (one being held in reserve), supplying steam 
for a cross-compound condensing air compressor of standard 
make. The steam cylinders have Meyer valve gear and are 16'' 
and 2^" diameter by 24'' stroke. The two-stage air cylinders are 
28'' and 18'' by 24". From a two weeks' run the following results 
were obtained. 

Total coal burned, lbs 264,300 

Total feed water, cu. ft 37)45Q 

Total feed water, lbs 2,335,568 

Average temperature of feed water, degrees Fah 131 

Average evaporation per lb. coal consumed, lbs 8.72 

Average revolutions per minute 63.1 

Indicated horse-power of steam end, corresponding to 63.1 

R.P.M. steam 161 

Corresponding indicated horse-power of air end 123 

Average steam pressure, lbs 115 

Average vacuum, lbs 10.5 

Average air pressure, lbs 96 

Average temperature of outside air, degrees Fah ,. . 24 

Average air piston displacement at 70° F., cu. ft 11 72 

Average metered output corrected to 70° F., cu. ft 758 

The average evaporation, of 8.72 lbs. of water from 131° F. 
to an average steam pressure of 115 lbs., is equivalent to 9.83 lbs. 
of water evaporated from and at 212° F. per lb. coal consumed. 
At the average compressor speed of 63.1 revolutions per minute, 
the metered output was equivalent to 758 cubic feet of free air per 
minute, the piston displacement being 1,172 cu. ft. per minute. 
Table VIII and Fig. 90 present the principal data of the test run on 
this compressor. 

To find the average operating results, the curves at 63 revolu- 
tions should be followed, at which the indicated H.P. of the steam 
cylinder was 160.8, and that of the air cylinder, 123, showing the 
mechanical efhciency to have been 76.5 per cent. The theoretical 
12 



178 



COMPRESSED AIR PLANT 



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PERFORMANCE OF AIR COMPRESSORS 



179 



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i8o 



COMPRESSED AIR PLANT 



H.P. required to compress isothermally one cubic foot of free air 
per minute to 96 lbs. (the average pressure) is 0.129. The 
theoretical useful work done by the compressor is, therefore, 



100 



90 



2 

I 70 



t 60 



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Fig. 92. — Compressor Plant No. i. 

758 X .129 or 97.8, and the net total efficiency of the compressor 
is 97.8/ 161 or 60.8 per cent. 

Table IX shows the actual cost of running this compressor 



PERFORMANCE OF AIR COMPRESSORS 



l8l 



at different speeds. The data were furnished by the owner and 
are based on one year's operation. In Fig. 91 these costs are 
plotted, showing how the cost per steam H.P. per year is affected 



80 

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1 — 

















20 40 60 80 100 120 

R.P.M. 

Fig. 93. — Compressor Plant No. 2. 



by the average running speed of the compressor. The curve of 
Fig. 92 shows the operating costs in another way. These 
costs may be read in terms of i,ooo cu. ft. of free air compressed 



l82 



COMPRESSED AIR PLANT 



to loo lbs. or i,ooo cu. ft. of compressed air at loo lbs. gauge 
pressure. 

Test of Plant Number Two. The plant consisted of three 
150 horse-power, return tubular boilers, supplying steam for a 





r' 






























































































































































































































































































































































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Curve A CoRl 

" B " and Labor 

« C " Labor andint, and Dep. 

" D " "■ Int. and Dep. and Supplies 


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10 



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90 



100 



Fig. 94. — Compressor Plant No. 2. 

Corliss engine, the air compressor, and steam heating. To deter- 
mine the boiler horse-power, a meter was placed on the steam pipe 
to the compressor during the test run, so that only the portion of 
steam actually used by the compressor was charged to the same. 
The compressor was duplex, with Meyer valve gear, simple steam 



PERFORMANCE OF AIR COMPRESSORS 



183 



cylinders 14'' X 22", and two-stage air end, 14'' and 22'' X 22'' 
stroke, rated by the manufacturer at 1,050 cubic feet of free air per 















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56 



10 20 30 40 50 60 70 80 90 100 

R. P. M. 

Fig. 95. — Compressor Plant No. 2. 

minute at 105 revolutions. At this plant the test lasted over a 
month, with the following results: 

Total coal consumed, lbs 459>25o 

Total feed water, lbs 2,496,000 

Average evaporation per lb. coal consumed, lbs 5-46 

Average revolutions per minute 3"- 

Corresponding average indicated horse-power (from curve) 53- 



184 COMPRESSED AIR PLANT 

Hourly readings of the revolution counter were taken, showing 
an average speed of 36.05 revolutions. At this speed the steam, 
consumption was 51 pounds per I.H.P. hour, as measured at the 
throttle, the air meter showing a delivery of 275 cubic feet of free 
air per minute. The total efhciency was 67 per cent. Taking 
the ordinary method of computing the mechanical efficiency only 
at the same speed, there would be 48 air horse-power, divided by 
54 steam horse-power, giving an efficiency of 89 per cent. 

The coal consumption per indicated horse-power per yeaf, as 
shown by the books of the company, amounted at the average 
speed to about 56 tons. Table X, with Figs. 93, 94, and 95, 
present the details of the test on this plant, which was con- 
ducted in a manner similar to that on plant No. i. 

Test of Plant Number Three, This plant consisted of two 125 
horse-power return tubular boilers, supplying steam for a non- 
condensing cross-compound air compressor of standard make; 
steam cylinders 18'' and 35''X24'', air cylinders 14" and 28''X24''. 
A two weeks' run gave the following results: 

Total coal burned, lbs 221,190 

Total feed water, cu. ft 34)273 

Total feed water, lbs 2,094,657 

Average temperature feed water, degrees Fah 154 

Average evaporation per lb. coal consumed, lbs 9.48 

Average boiler horse-power 208 

Average revolutions per minute i . 66 

Average indicated horse-power of steam end, at 66 R.P.M (from 

curve) 210 

Average indicated horse-power of air end (from curve) 128.5 

Average steam pressure 97 

Average air pressure . 97 

Average outside temperature, degrees Fah 23 

Average air piston displacement at normal speed, cu. ft., at 70° F. 1,372 

Metered output in cu. ft. corrected to 70° F 734 

The average evaporation of 9.48 lbs. of water per pound of 
coal, from 154° F. to an average steam pressure of 97 pounds, is 
equivalent to 10.4 pounds of water evaporated from and at 212° F. 
At the average speed of 66 revolutions, the displacement was 1,240 
cubic feet of free air per minute, while the metered output was 
734 cubic feet, showing a net volumetric efficiency of 59 per cent. 



PERFORMANCE OF AIR COMPRESSORS 



l8: 



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i86 



COMPRESSED AIR PLANT 



To determine the conditions in average operation, the curve at 
66 revolutions should be followed (Fig. 96), at which the indicated 
horse-power of the steam cylinders was 210, and that of the air 
cylinders, 128. This shows the efficiency to be 61 per cent., the 
friction loss being 81.5 horse-power, or 39% of that delivered by 
steam end. This extremely high friction loss was due to the fact 
that the compressor shaft was out of line, and the plant could not 







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Fig. 96, — Compressor Plant No. 3. 



be shut down long enough to rectify it. The details and results 
of this test, given in Table XI and Figs. 96, 97 and 98, are interest- 
ing in exhibiting the inefficiency that may be caused by a purely 
mechanical defect. 

Test Number Four. The results of a test on another plant 
are given in Table XII and Fig. 99, the details of the boiler 
test and of the costs being omitted. In this case the compressor 
was of the tandem compound non-condensing type, with Corliss 



PERFORMANCE OF AIR COMPRESSORS 



187 



valve gear for the steam cylinders. The test shows that, at a low 
speed, the steam consumption increases more rapidly than vrith 
the Meyer type of valve. 

600 



550 



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Fig. 9 7. ^Compressor Plant No. 3. 

Summary. The results of these tests are enlightening, in 
showing the actual amount of the losses occurring in the com- 
pression of air, particularly when the compressor is operating 
under the unfavorable conditions of varying air consumption, 



COMPRESSED AIR PLANT 



necessarily obtaining in mining and other work in which machine 
drills play an important part. These losses are always recognized 
as existing, by compressor builders and by intelligent users, and J 



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Fig. 98. — Compressor Plant No. 3. 



O 

o 



it is clearly desirable that properly conducted tests should be made 
more frequently. 

Again, compressor plants generally develop less power than 
their full rated capacity. It should be remembered that an air 



PERFORMANCE OF AIR COMPRESSORS 1 09 

compressor is essentially a variable speed machine, its speed being 
regulated by some form of throttling governor, connected with the 
air-pressure regulator. The machine is therefore called on to 
run only as fast as the demand for air may require. It may be 
suggested that it would be well for compressor builders to give in 







Ann 








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Fig. 99. — Compressor Plant No. 4. 



their catalogues the actual horse-power rating at different speeds, 
with a table of efficiencies at different loads and speeds, just as is 
done by some of the manufacturers of electrical machinery. 
Catalogues might also include some definite data respecting the 
cost per horse-power delivered by the air end of the compressor at 
different working speeds. 



CHAPTER XI 

AIR RECEIVERS 

On being discharged from the compressor cylinder the air is led 
into a receiver before passing to the air main. Users of compressed 
air have been slow to realize the important part played by the air 
receiver in the economical operation of compressors, and until 
recently insufficient attention has usually been given to questions 
of its capacity, design, and position relative to the compressor. 

In its common form the receiver consists merely of a cylindrical 
shell of steel plate, resembling a steam boiler v^ithout tubes or 
flues. It is provided with pipe connections to the compressor and 
air main, a pressure gauge, safety-valve, drain cock, and man-hole. 
The receiver may be set vertically or horizontally, the vertical form 
being generally preferable, as it occupies less floor space (Fig. loo). 
Another design, which may also be employed as an intercooler, is 
illustrated in Fig. 56. The cubic capacity of the receiver 
should be properly proportioned to the size of the compressor. 
The dimensions range from, say, 24 ins., diameter by 4 or 6 ft. 
long, up to 48 or 60 ins. by 14, 16, or 18 ft., the largest sizes hav- 
ing a capacity of from 200 to nearly 400 cu. ft. Receivers are 
usually built to stand a test of 165 lbs. cold-water pressure, for 
working under pressure of ico to 120 lbs., higher pressures than 
this being rarely necessary in ordinary practice, such as mine 
service. The shells are single-riveted on circular seams and, ex- 
cept for small sizes, double-riveted on longitudinal seams; the 
heads being dished or hemispherical. To produce the best results, 
the receiver should be placed close to the compressor, or in any 
case not more than 40 to 50 ft. distant. A large horizontal re- 
ceiver is shown in Fig. 10 1. 

The principal functions of an air receiver may be summarized 

190 



AIR RECEIVERS 



191 



as follows: (i) to eliminate the 
pulsating effect of the strokes of 
the compressor piston and pre- 
vent rapid fluctuations of press- 
ure; (2) to minimize the fric- 
tional loss attending the flow of 
air through the lines of piping; 

(3) to serve in some degree as an 
equalizer and reservoir of power ; 

(4) to cool the air as thoroughly 
as possible before it passes into 
the main, thus causing it to de- 
posit a part of its moisture in the 
receiver, whence it is drained off. 

Regarding the first point, the 
volume of the receiver should be 
sufficiently great in proportion 
to that of the compressor cylin- 
der to prevent any material rise 
of pressure in the receiver by the 
incoming volume of air forced 
into it at each stroke. If the 
compressed air were discharged 
directly into the main, large 
fluctuations of pressure would 
occur, accompanied by peri- 
odic acceleration of flow. This 
would not only increase the fric- 
tional resistance in the pipe, but 
at the end of each stroke the 
compressor piston would have to 
force the air out of the cylinder 
against a pressure momentarily 
greater than the normal. A 
loss of power would thus be 
caused, and the variation of the 




Fig. 100. — Vertical Air Receiver (Nor- 
walk Iron Works Co.). 



3-g2 COMPRESSED AIR PLANT 

work done throughout the stroke of the piston would be increased. 
The violence of the discharge pulsations is obviously greater with 
a single cylinder than a stage compressor, working to the same 
air pressure, because the total discharge must take place from the 
cylinder of larger diameter in a smaller proportion of the length of 
stroke than is the case with the high-pressure cylinder of a stage 
compressor. In the latter the delivery valves open earlier in the 
stroke, and the air pipe is about one-half the diameter of the cylinder. 



12 AIR OUTLET-; 



^^J 


^frn^Mrn^^ 


1 -^ 


^^^a_^ 


4 WATtR J_ 
OUTLET|T§=f 


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jTiG. I o I. —Horizontal Receiver-Aftercooler (Ingersoll-Rand Co.). 

The second function of the receiver is best fulfilled by placing 
an auxiliary receiver near the point at which the compressed air is 
used. Just as the receiver at the compressor diminishes the momen- 
tary rise of pressure in the main caused by each stroke of the pis- 
ton, so a second receiver close to the engine or machine using the 
air will prevent a drop of pressure as each cylinderful of air is 
drawn off. By reducing the fluctuations of pressure the two re- 
ceivers maintain a practically constant flow of air through the main 
connecting them and the friction and loss of pressure are thus 
minimized.* 

■ For mine service the second receiver would usually be placed 
somewhere underground. This arrangement is always advan- 
tageous when the air main is of great length. Underground re- 
ceivers are not often used for air drills alone, but they become a 

* Other questions relating to the flow of air in pipes, frictional losses, etc., are 
discussed in detail in Chapter XVI. 



AIR RECEIVERS 



193 



necessity when large machines, such as pumps and hoists, are run 
by compressed air. They are useful, moreover, in permitting a 
further deposition of moisture from the air, thus rendering the air 
dryer and more suitable for use in expansive-working engines. To 
be most effectual in accomplishing this, the underground receiver 
should be placed at the point in the pipe line where the air has 
reached its lowest temperature — a consideration not always con- 
sistent with the local conditions. 

Undergroimd receivers are usually similar in construction to 
those installed near the compressor. Sometimes, however, as for 
example at the Mansfeld copper mines, Germany, another mode of 
construction has been satisfactorily adopted. A chamber is ex- 
cavated in the rock, all loose stone removed, and the walls cemented 
tight. The chamber is closed by a brick dam composed of two 
parallel walls, with a two-inch layer of cement between them. In 
the dam are set a cast-iron man-hole with suitable cover, several 
pipes for connecting with mains to the various working places, and 
a drain pipe and cock close to the floor. The latter is opened from 
time to time, to blow out the accumulated water and sediment. 
A pressure gauge is attached to the man-hole cover. Such reser- 
voirs may be built to cost much less (for large sizes) than ordinary 
shell receivers of equal capacity.* 

The third function of the receiver is apt to be misunderstood 
or exaggerated. While it is true that it acts to a limited extent as 
a reservoir of power; yet, to be of much practical service in this 
respect, its capacity must be very large. 

For example, take a 2C-in. compressor, working at 60 lbs. 
pressure to supply air for a regular consumption. To enable the 
receiver to meet the demand for only i minute after the com- 
pressor is stopped, and not have the pressure fall more than 15 lbs., 
it would have to be 5 ft. diameter by 50 ft. long. Again, if the 
compressor were running at a constant speed and the demand for 
air should suddenly increase 25 per cent. — as might happen in 

* Zeitschrijt fiir das Berg-, Hiitten-und Salinen-Wesen, Vol. XLI, p. 119. A 
receiver of the kind mentioned was built at Mansfeld for about one -third the cost of 
an equivalent steel receiver. 



194 COMPRESSED AIR PLANT 

starting several more machine drills — a receiver of the size men- 
tioned could meet the extra demand only 4 minutes.* It is thus 
evident that while a receiver is useful as an equalizer within cer- 
tain limits, yet, unless it be large, the pressure might quickly run 
up to an unreasonable amount in case of an unexpected decrease 
in consumption of air. Long pipes of large diameter assist in 
equalizing the flow of air, but their use does not preclude the 
necessity of receivers. It is much cheaper to employ piping of 
moderate size, in connection with a receiver of generous di- 
mensions. 

Probably the most important ofhce of the receiver is to cool 
the air before it passes into the main. In recent years much more 
attention than formerly has been given to this point. The velocity 
of flow of the air coming from the compressor is greatly reduced on 
entering the relatively large volume of the receiver; it is cooled 
somewhat at the same time, and caused to deposit a part of the 
moisture in suspension, which otherwise would be conveyed into the 
system of piping, and thence to the machines using the air. It is 
intended that the receiver shall be of sufficient capacity to drain the 
air as thoroughly as is economically practicable. But in the or- 
dinary sizes of shell receiver the results are usually quite imperfect,, 
because the air passes through too rapidly to permit any large drop 
in temperature. The inlet and outlet pipes of the receiver should 
be placed in proper relative positions. If at opposite ends, and 
especially if these pipes point toward each other, a strong through 
current is caused, which reduces the usefulness of the receiver. A 
large part of the entering volume of air passes out again without 
having had time to cool or to drop much of its entrained moisture. 
One mode of arranging the pipe connections is to place the inlet on 
one side, near the end of the receiver, while the outlet is at the 
opposite end, in the middle of the head. The air is thus forced to 
change its direction of flow. Or, as in Fig. 100, both pipes may be 
connected near the top, the outlet pipe being carried through the 
receiver nearly to the bottom, where the air is likely to be slightly 
cooler (and dryer) than at the top. As the inlet pipe shown in this 

* Norwalk Iron Works Catalogue, 1906, p. 63. 



AIR RECEIVERS I95 

case is connected tangentially to the periphery of the receiver, a 
rotary motion is imparted to the body of air, so that each particle 
remains longer in the receiver and under its cooling influence. 
Some receivers are provided with baffle-plates for the same pur- 
pose, as in Fig. loi. With wet compressors a large amount of 
moisture is carried into the receiver; even in dry compressors some 
water collects from the natural moisture of the atmosphere, espe- 
cially in warm weather. Part of the lubricating oil carried over from 
the compressor cylinder is also deposited in the receiver. At inter- 
vals, according to atmospheric and other conditions, the water 
and oil are drained off by means of the cock provided for the 
purpose. 

Another result of cooling in the receiver may be noted. A 
receiver of ample size, placed close to the compressor, tends in 
some degree to economize power; because, whatever cooling is 
accomplished reduces proportionately the temporary increase of 
pressure due to the heat of compression. Hence, the piston con- 
sumes somewhat less powTr in forcing the air out of the cylinder 
against the receiver pressure than if the air were left to cool 
gradually in a long length of piping. As the heat of compression 
must be lost in any case before the air is used, this saving is worth 
while, however small it may be, since it is produced without cost and 
incidentally to the normal operation of the receiver. 

This consideration has of late led to the employment of what 
are called "receiver after-coolers," (Fig. loi), practically identical 
in construction with the large tubular intercoolers shown in Figs. 
54 and 56.* The shell contains a series of water-cooled tubes, 
between which the air is caused to circulate before passing to the 
larger outer portion of the receiver, whence it is discharged into 
the main. Having a sufficient volumetric capacity and cooling 
area of tubes, this type of receiver cannot fail to he more efficient 
as an after-cooler and the benefits of employing a receiver are more 
fully realized. 

* These are referred to, in the latter part of Chapter VI, as being applicable as 
intercoolers for stage compressors. See also an article by Frank Richards, in 
Compressed Air, Jan., 1907, p. 4329. 



CHAPTER XII 

SPEED AND PRESSURE REGULATORS FOR COM- 

PRESSORS 

If the consumption of compressed air were constant, no more 
regulation of the compressor's speed and power would be required 
than that furnished by an ordinary governor for the steam end, to 
take care of fluctuations in boiler pressure or accident to some part 
of the mechanism. But the conditions under which most air com- 
pressors operate make it necessary to provide for running econom- 
ically even when there are wide variations in the rate at which the air 
is used. In event of a sudden temporary decrease in consump- 
tion, the compressor must be slowed down, the alternative being 
to blow off air at the receiver safety valve, just as steam would be 
blown off in similar circumstances from a boiler. As a cubic 
foot of compressed air, however, costs more than a cubic foot of 
steam, the air cannot be allowed to go to waste at a safety valve. 
The compressor must be furnished with some device for coor- 
dinating the quantity of steam admitted to the steam end with the 
variable air pressure in the receiver, thereby regulating the piston 
speed in accordance with the demands upon the air end. Further- 
more, it is not enough to provide only for varying the speed of the 
compressor. At times, the consumption of air may cease entirely 
for a short period, and, to avoid the necessity of bringing the com- 
pressor to a standstill, provision should be made for unloading the 
air end. When this is done useful work stops for the time being, 
the compressor consuming only enough steam to overcome friction 
of the moving parts, and turn its centers. 

Numerous regulating and unloading mechanisms have been 
devised, so that instead of requiring the almost constant attendance 
of an engineer at the throttle, the modern air compressor operates 

196 



SPEED AND PRESSURE REGULATORS FOR COMPRESSORS 1 97 

automatically under the widest variations of load. As these useful 
devices differ greatly in design, the subject will best be illustrated 
by giving a few examples in detail. They may be classified under 
two heads: (i) speed governors and pressure regulators; (2) un- 
loaders for the air cylinders. 

Speed Governors and Pressure Regulators. Speed governors 
are usually of the ordinary centrifugal or fly-ball type, and may be 




Fig. 102. — Clayton Governor and Pressure Regulator. 



applied to the steam end of the compressor merely to regulate its 
speed, as in case of a steam engine; or their action may be so modi- 
fied and controlled by the changing receiver pressure as to produce 
a combined speed and pressure regulation. The air cylinder is not 
completely unloaded at any time, the compressor being simply 



198 COMPRESSED AIR PLANT 

speeded up or slowed down in conformity with the rate at which 
the air is used. 

The pressure regulator of the fly-ball governor type may be 
illustrated by Fig. 102 (Clayton governor) . The stem of the throttle 
valve, h, which is inserted in the steam pipe, connects with the 
spindle of the ball governor, by which the speed of the compressor 
is limited and controlled. At p is shown the bevel gearing for 
operating the governor, a small pulley being mounted on the gear 
shaft and driven by belt from the crank-shaft of the compressor. 
By means of the weighted lever, i, and the small air cylinder, y, the 
action of the ball governor is modified by the air pressure in the 
receiver. Air from the receiver enters the cylinder, y, through 
the pipe, k, and when the pressure exceeds its assigned limit, raises 
the piston and weight, and shuts off steam by forcing down the 
throttle valve, h, the pressure of the lever being applied at the point, 
/. The governor may be adjusted to its work by the spring and 
thumb-screw, m, acting on the small lever, n, which tends to keep 
open the throttle against the downward pressure of the weighted 
lever, z, upon the valve stem. The spring, 0, is introduced to ease 
the drop of the weight when the air pressure falls. 

Other designs, similar in general principle but varying in 
many details, are used on the Ingersoll-Rand, Sullivan, Franklin, 
McKiernan, American, and other compressors, when steam- 
driven and of the straight-line or duplex type. The Sullivan 
speed and pressure regulator, as supplemented by an unloading 
attachment, is described hereafter. 

An entirely different form of governor is the Norwalk (Fig. 
103). A balanced throttle valve, a, is placed in the main steam 
pipe, and above it is set a small air cylinder, h, the piston rod of 
which is a prolongation of the valve stem, c. At the side of the 
cylinder, h, is a spring safety valve, d, connected by a pipe, e, with 
the receiver, or with the air main leading to it. By means of a 
hand-wheel, /, on the safety valve, the spring is adjusted so. that the 
air will lift the valve, and pass through it, at any desired pressure. 
When the receiver pressure exceeds this limit the safety valve, d, 
rises and allows air to pass under the piston in the small cylinder, 



I 



SPEED AND PRESSURE REGULATORS FOR COMPRESSORS 1 99 

h, raising it and partly closing the throttle. If no escape were 
provided the piston would be forced at once to the top of the cylin- 
der. To regulate its movement and prevent shutting off the 
steam completely, a very narrow vertical slot is cut in the side of 
the cylinder. As the piston rises, more and more of this slot is un- 
covered and furnishes an escape for the air passing into the cylin^ 




Fig. 103. — Norwalk Pressure Regulator. 



der. The slot being very narrow, a slight difference in the quan- 
tity of air causes the piston to assume a high or low position. In 
this way the throttle is moved, controlling the admission of steam 
and the compressor speed. As the air pressure falls the valve 
begins to open again. To prevent the small piston from rising 
too far and stopping the compressor by completely closing the 
throttle, a screw stop, g, is set in the top of the regulating cylinder, 



200 



COMPRESSED AIR PLANT 



b. This can be so adjusted by hand that, when the small piston 
has reached the top of its stroke, just enough steam is admitted 
by the throttle to keep the compressor in motion. 

In another form of this governor, shown in vertical section, Fig. 
104, the fine slot in the little cylinder, B, is replaced by a tapered 




Fig. 104. — Norwalk Pressure Regulator. 

recess in the stem or piston rod of this cylinder, at the point where 
it passes through the lower head (indicated at R, in the small cut 
to left of main figure). As the piston in this cylinder is forced 
upward by the air pressure the area of the opening formed by 
the slotted stem furnishes a graduated escape for the air, and so 



SPEED AND PRESSURE REGULATORS FOR COMPRESSORS 20I 

regulates the small piston's movement, and through it the throttle 
valve. 

The IngersoU-Rand Co. also makes a steam regulator in which 
the stem of the main throttle is prolonged to the piston of a small 
horizontal air cylinder, attached to the side of the throttle. This 
piston is moved by air pressure conveyed through a quarter-inch 
pipe from the receiver. 

Another governor (Clayton) is shown in Fig 105. The 
throttle valve, a, is provided with a lever and weight, b, connected 




Fig. 105. — Clayton Pressure Regulator. 

with the valve stem, c. Close alongside of the throttle, and for 
convenience clamped on the steam pipe, d, is a small air cylinder, 
e, the upper end of whose piston rod is pinned at / to the weighted 
lever. Entering the bottom of this air cylinder is a small pipe, 
g, from the air receiver. When the pressure in the receiver ex- 
ceeds the assigned limit the weighted lever is raised and, partially 
or wholly, closes the steam throttle, a. Then, when the air press- 



202 COMPRESSED AIR PLA.NT 

ure has been reduced by the slowing down of the compressor, and 
by consumption of air from the receiver, the weight falls and re- 
opens the throttle. 

With governors and regulators of the type represented by the 
preceding examples, the operation and control of the compressor 
is not automatic under all conditions, but they answer the pur- 
pose for some kinds of service. In case no air is drawn from the 
receiver, the compressor slows down until it just passes its 
centers; then, in most of the designs, if the pressure continues to 
rise, a little air will blow off at the receiver safety-valve, or the 
compressor may be stopped completely by closing the steam 
throttle. 

Air-Cylinder Unloaders. These are designed to exercise com- 
plete automatic control over the compressor, when the latter is 
belt-driven; and also for steam-driven compressors when used in 
conjunction with a governor. The throttle is first nearly closed 
(in steam compressors), as the consumption of air decreases; 
then, if it ceases altogether, the unloading mechanism either shuts 
off the intake air or else opens the discharge valves, thus ad- 
mitting air at receiver pressure to both ends of the cylinder. In 
either case the pressures on opposite sides of the piston are bal- 
anced and all useful work ceases, though the compressor contin- 
ues to turn its centers, taking only enough steam to overcome 
friction. 

The Rand " Imperial " unloader, for small compressors 
driven by belt or direct-connected electric motor, furnishes an 
example of this type of regulator (Fig. io6). It is inserted in the 
intake pipe, and shuts off the air from the inlet valves when the 
receiver pressure rises above the set limit. In the cut the passage 
of the intake air is shown by the arrows. The small chamber 
(6g) is connected by a J-in. pipe with the receiver. As the 
pressure increases, the piston (57) moves against the resistance 
of the spring (56), admitting receiver air, through the small ports 
on the left of the piston, to the lower side of the plunger valve 
(61). On reaching its seat this plunger closes the intake to the 
compressor cylinder. The resistance of the spring (56) miay be 



SPEED AND PRESSURE REGULATORS FOR COMPRESSORS 203 

adjusted by the screw-plug (55), for any required working press- 
ure. As the receiver pressure falls again, on increased con- 
sumption of air, the spring forces down the piston (57). This 
closes the lower small air port, leading to the. under side of the 




SECTIONAL VIEW 
RAND IMPERIAL UNLOADER 



Fig. 106. 



K Pipe 
from Receiver 



plunger valve (6i), and at the same time opens the upper hori- 
zontal port, connecting with the open screw-plug (55). The 
air below the plunger valve is thus exhausted, causing the latter 
to fall, thus reopening the intake passage. The useful work of the 
compressor is then resumed. 



204 COMPRESSED AIR PLANT 

An unloader similar to the above is used for some of the Allis- 
Chalmers compressors. The Ingersoll-Rand Co. makes an 
automatic '' choking " controller, which is applied to the intake pipe 
of the piston-inlet compressor. It is adjustable for any desired 
limit of pressure by a weighted lever, and may be used for all 
forms of steam-driven compressors. 

A recent form of this unloader is shown on the folding plate, 
Fig. 33, as applied to one of the latest designs of the Ingersoll- 
Rand compressors — ''Imperial," type lo. The same company 
is now making a clearance or expansion controller, for unloading 
the air end of the compressor. It varies the clearance volume of 
the cylinder by automatically varying the number of discharge 
valves in action. A small air cylinder is placed between the main 
cylinders and connected with the receiver. As the receiver 
pressure increases, the piston of the controller cylinder rises higher 
and higher. Inserted in the side of this cylinder is a series of 
small pipes, each connected by branches with a discharge valve 
on opposite ends of the main cylinder. These valves are thus 
released from the receiver pressure successively, as the pressure 
increases, and the work done by the compressor is proportionately 
reduced. 

A combined governor and pressure regulator, with unloading 
attachment, as employed by the Sullivan Machinery Co., will 
illustrate a type of compressor regulator that has been adopted by 
several builders, though with many variations in details of design 
(Fig. 107). It may be used with straight-line or duplex, steahi- 
driven compressors. The split-ball governor (11), belt-driven 
from the crank-shaft to the pulley (20), accompanied by the 
tightener (43), controls the steam throttle (3). Connected 
with the governor spindle and throttle valve stem, at 28, is a 
lever (25), the position of which is influenced by the centripe- 
tal action of the set of springs (31,32, and 26). By screwing 
up or down the hand-wheel and speeder screw (5), this system of 
springs (and with them the governor) is set to run the compressor 
at any desired speed. The other element of the governor is the 
air-pressure device, which, by the position of the plunger in the 



SPEED AND PRESSURE REGULATORS FOR COMPRESSORS 205 

small air cylinder (i8), causes the springs to be brought into action 
in the order of their strength, thus producing movement of the 
lever (25). 

The pressure device is connected with the air receiver by the 
union valve (33), admitting air to the little cylinder (27), the piston 



Oil Oov. B«re 




rm 



Fig. 107. — Sullivan Governor and Unloader. 



of which operates a needle valve. This valve is held closed 
against any desired minimum air pressure by the adjustable 
weight (36) and the regulating screw and spring (37 and ^S). 
When the receiver pressure rises above the normal, it opens the 
needle valve and admits receiver air to the cylinder (18). As the 
pressure increases, the plunger in (18) rises against the counter- 
spring (26) and through the lever (25) tends to close the main 
steam throttle (3), thus slowing the compressor. Total stoppage 



2o6 



COMPRESSED AIR PLANT 



is prevented by screwing down the nut of the stop-screw (23), so 
as to limit the upward movement of the pressure plunger. This 
plunger is designed to act intensively, being so proportioned that 
a -variation of only 2 or 3 lbs. receiver pressure is multi- 
plied to 40 lbs. in its action on the governor. In this way a sen- 
sitive control is produced within narrow limits of working air 
pressure. To prevent violent movements of the pressure ele- 




Connection with 
. Discharge Valve 

Fig. 108. — Ingersoll-Sergeant Regulator and Unloader. 

ment, in case of sudden changes of receiver pressure, the 
plunger in (18) is provided with an oil dash-pot. 

A somewhat similar pressure regulator and unloader is used 
on the Franklin compressor.* 

Another unloader, applicable to straight-line and duplex 
compressors, and in a modified form to stage compressors also, is 
the Ingersoll-Sergeant. It differs materially from the unloaders 
previously described, in controlling the action of the discharge, 
instead of the inlet valves. The principles of its construction 
and operation will be understood by reference to the longitu- 

* Mines and Minerals, May, 1905, p. 504. 



SPEED AND PRESSURE REGULATORS FOR COMPRESSORS 207 

dinal section in Fig. 108. The most recent design of this unloader 
differs in some details from that shown in the cut, but its mode 
of working is unchanged, and many are in use. 

A weighted plunger, a, working in a small cylinder, is at- 
tached for convenience to the shell of the compressor cylinder. 
From the chest in which a is set there are four pipe connec- 
tions as shown : h leads to a balanced throttle valve in the 
main steam pipe, c connects with the air receiver, and d and 
e with one or more discharge valves at each end of the cylin- 
der. The stem of the steam throttle, /, is a continuation of 
the piston rod of a small horizontal air cylinder, g, which is 
attached to the side of /. Behind the piston of this little cylin- 
der enters the air pipe, h. When the pressure in the receiver 
becomes too great the safety valve, a, rises, and exhausts the air 
behind the two discharge valves which are connected with the 
pipes, d and e. This admits air at receiver pressure into each end 
of the compressor cylinder, thus balancing the pressure on the 
two sides of the piston and unloading the engine. At the same 
time the air in the little cylinder, g, is also exhausted, so that the 
throttle valve, /, moves to the right and admits only enough steam 
to keep the compressor slowly turning. When the compressor is 
thus unloaded no work is done; the air is merely circulated through 
the pipes, d and e, from one end of the cylinder to the other, until 
more air is drawn from the receiver and the pressure reduced. 
Then the safety valve, a, closes and the pipes, d and e, are again 
filled with compressed air. The steam throttle is also forced open 
by the pressure through the pipe, 6, and compression goes on 
regularly. The admission and discharge lines of an air card from 
a compressor thus unloaded form practically a single horizontal 
line, at a height above the atmospheric line representing the 
receiver pressure. 

For steam-driven compressors of the Corliss type, as built by 
the Ingersoll-Rand, Nordberg, Laidlaw-Dunn-Gordon, Sulli- 
van, Alli^-Chalmers, and some other companies, the air-press- 
ure regulators act in conjunction with ball or other centrifugal 
governors. All of them control the operation of the compressor 



208 



COMPRESSED AIR PLANT 



by acting upon the expansion gear of the steam end and chang- 
ing the point of cut-off. 

The Laidlaw- Dunn- Gordon governor (Fig. 109) maybe taken 
as an example. Air is admitted from the receiver to the small 
cylinder, A, the piston of which is weighted, as shown. The 




Fig. 109. — Laidlaw-Dunn-Gordon Air Governor. 



SPEED AND PRESSURE REGULATORS FOR COMPRESSORS 209 

action of the lever, B, supporting the weight is adjusted by the 
coil spring, C. This lever is linked to a floating lever, D, pinned 
to the vertical side rods of the ball governor. D is connected by 
the link E to the bell-crank, F, the lower arm of which is pin- 
connected to the long horizontal rod, G. By this system of levers, 
the movement of G, and through it the point of cut-off of the Cor- 
liss steam gear, is under the combined control of both ball gov- 
ernor and of the receiver pressure as influencing the position of 
the piston of the cylinder A. The arm, H, is pivoted at the foot of 
the governor post. Connected to it are the cam, I, and the idler 
pulley, J, which rests on the governor belt. In case the belt 
breaks, the idler pulley falls and the cam allows the governor 
to drop, thus shutting off steam and preventing the compressor 
from racing. The designs of governors of this type are worked 
out in a number of different ways. 

iVnother example of governor is that employed on the constant- 
speed, variable-delivery compressor, built by the Nordberg 
INIanufacturing Co. It is for m.otor-driven machines, with 
Corliss air valves, and operates by closing the inlet valve before the 
stroke is completed.* During the remainder of the forward 
stroke, the air already admitted to the cylinder is expanded below 
atmospheric pressure, and is then compressed on the return 
stroke. This is practically equivalent to varying the working 
length of the stroke. 

A general view of this compressor is shown in Fig. no, and 
the construction of the valve gear is indicated in Fig. in. 
The wrist-plate w is driven by the rod a, from an eccentric on 
the fly-v/heel shaft; another eccentric drives the releasing mech- 
anism, through the rod b, which osciflates the arm c about the fixed 
center d. Swivelled to the lower end of c is a 3-armed rocker. 
The arm i is linked by the rod j to the radius fork k, which in 
turn is connected to the pressure governor /. The arms g and h 
of the rocker, through the rods e and/, operate the releasing cams 
n and 0, which are attached to the forward and back inlet-valve 
spindles. When the compressor is working regularly, under 

* American Machinist, Aug. 22d, 1907. 



SPEED AND PRESSURE REGULATORS FOR COMPRESSORS 211 




a, 

S 
o 
U 




Fig. 112. — Detail of Valve Gear Shown in Fig. m. 



SPEED AND PRESSURE REGULATORS FOR COMPRESSORS 213 

normal consumption of air, the rocker arms g and h maintain a 
vertical position, under the action of the eccentric rod b and 
impart equal movement to both knock-off cams. If, however, 
the receiver pressure increases, the rocker arm i will be drawn 
upward, the arms g and h will take an inclined position and, 
through the rods e and/, the point of release of the A^alves is altered. 

The releasing mechanism is shown by Fig. 112. Mounted 
on the valve spindle is a rocker having three arms, a, b, and c. 
The wrist-plate link is connected to a, the releasing latch d to b, 
and the governor cam-arm e to c. Part e is connected also to 
the governor by the rod /, as explained above, and hence has a 
compound motion: it swings bodily about its swivel pin at the 
top, and its position is adjusted laterally by the action of the 
governor. The cam slot has two circular arcs, struck from the 
center at the upper end of e, with an inclined jog connecting them. 
Since the roller on the arm g swings about its center under the 
action of the cam groove, as the cam is moved from the main 
eccentric by the rod/, the latch d is alternately released and en- 
gaged, when the roller passes the jog in the cam. The point of 
the stroke at which the release takes place is determined by the 
governor, as already stated. 

In Figs. 113, 114 and 115, are given a set of indicator cards from 
a two-stage compressor provided with this regulating mechan- 
ism, and running at a speed of 74 revolutions per minute. 
The upper card in each cut is from the intake or low-pressure 
cylinder, the lower card from the high-pressure cylinder. Fig. 
113 shows the cards when working at nearly full load. Fig. 114 
(half load) illustrates the action of the regulating gear. Taking 
the crank-end card, C, the inlet valve remains open from the 
beginning of the stroke, at a, approximately to mid-stroke, b, 
at which point the releasing gear acts and the valve closes. From 
b to the end of the stroke, at c, the air in the cylinder expands 
below atmospheric pressure. On the return stroke, the com- 
pression line nearly coincides with the expansion line from c, until 
atmospheric pressure is reached at the point b, after which com- 
pression proceeds in the usual manner. The action of the inlet 



214 



COMPRESSED AIR PLANT 




SPEED AND PRESSURE REGULATORS FOR COMPRESSORS 215 

valves of the high-pressure cylinder is the same, except that the 
expansion and re-compression of the air is from receiver pressure, 
instead of atmospheric pressure. In Fig. 115 the cards show the 
very small amount of work done when the compressor is operating 
under nearly zero load. To simplify the mechanism, each cylinder 
is provided with its own pressure governor. 



chapte:r XIII 

AIR COMPRESSION AT ALTITUDES ABOVE SEA- 
LEVEL 

Because of the diminished density of the atmosphere, air 
compressors do not produce the same results at high ahitudes 
as at sea-level. Their effective capacity is reduced because a 
smaller weight of air is taken into the cylinder at each stroke. 
It is necessary, therefore, to modify the figures relating to the 
capacity and performance of compressors, as set forth in the 
first part of Chapter X. This matter is of special importance 
in connection with mining operations, because of the large num- 
ber of mines situated in elevated mountain regions. The rated 
capacities of compressors, in cubic feet of air, as given in the 
makers' catalogues, are for work at normal atmospheric pressure,, 
and due allowance must be made for decreased output at eleva- 
tions above sea-level. This reduction in output, which is usually 
also tabulated in handbooks and catalogues, should receive 
due consideration in order to avoid serious errors. For example, 
the volume of compressed air delivered at 60 lbs. pressure, at 
10,000 ft. elevation, is only 72.7 per cent, of the volume de- 
livered at the same pressure by the same compressor, at sea-leveL 
In other words, a compressor which at sea -level will supply power 
for 10 rock-drills, will at an elevation of 10,000 ft. furnish air 
for only 7 drills. 

The foregoing statement relates only to the volumetric capac- 
ity of the compressor. It must be remembered that the heat of 
compression increases with the ratio of the final absolute pressure 
to the initial absolute pressure. As this ratio increases with the 
altitude, more heat will be generated by compression to a given 
pressure at high altitudes than at sea-level. This additional heat 

216 



AIR COMPRESSION AT ALTITUDES ABOVE SEA-LEVEL 217 

temporarily increases the pressure of the air in the cylinder, while 
under compression, and more power is therefore required to com- 
press and deliver a given quantity of air. The corresponding loss 
of w^ork, due to the subsequent cooling of the air in receiver and 
piping, also increases with the altitude. 

Contrary to a common impression, the volume of air de- 
livered by a given compressor does not bear a constant ratio to the 




Fig. 



barometric pressure, but at different altitudes this volume de- 
creases slower than the barometric pressure. This relation may 
be shown as follows.* Two ideal indicator cards are represented 
in Fig. 116 one of a compressor working at sea-level, with an in- 
itial pressure Pi, the other at an altitude with a lower initial 
pressure Pj. The initial volume V and the final gauge pressure P, 
are the same for both compressors, P3 and P^ being the respective 
final absolute pressures. Vi and V2 are the final volumes, corre- 
sponding to the dotted isothermal curves, these volumes being 
taken as the basis, because they are those to which the com- 

* The general method of demonstration here given, together with Fig. ii6 and 
accompanying table, are taken by permission from an article by F. A. Halsey, in 
American Machinist, June 2d, 1898, p. 27. 



2l8 COMPRESSED AIR PLANT 

pressed air will eventually shrink on losing the heat of com- 
pression. From the theory of air compression, 

V ^P, 

V, P. 



VP, = V,P3, or ,- = ^ (I) 



andVP, = V,P„or^ = |^ (2) 

But since P^ = P^ + P, and P^ = P^ + P, equations (i) and 
(2) may be written: 

V P, + P P , , 

v:^-pr=^ + p: ^'^ 

1 V P, + P P 
and^=-^ — = ^ + p; (4) 

Dividing equation (3) by equation (4) : 
P^ 

y^ = ^p^,orV, :V, ::i + p-:i + p- (5) 

This gives an expression for the ratio between pressure and 
volume at sea-level and for any altitude above sea-level, of which 
the corresponding barometric pressure is P2. Thus, let P2 = 
10 lbs., P = 90 lbs., and Vi (from Table VII, page 165) =0.1404 cu. 
ft. By substituting these quantities in equation (5), V2 is found 
to be 0.0999, or approximately o.i cu. ft. 

In Table XIII, columns 4 and 5, are given the relative 
volumetric outputs, at gauge pressures of 70 and 90 lbs. 
of a compressor working at different altitudes, the figures being 
percentages of the normal output at sea-level. These per- 
centages have been derived by Mr. Halsey from equation (5), a 
constant loss of initial pressure of 0.75 lb. being assumed to 
allow for the resistance presented by the inlet valves, to which 
reference has been made in another chapter. That is, for 
practical purposes the sea-level atmospheric pressure is taken as 
14, instead of 14.7 lbs. The other columns show the mean 
effective pressures and indicated horse-powers, corresponding to 
different altitudes, up to 15,000 ft., which will be found con- 



AIR COMPRESSION AT ALTITUDES ABOVE SEA-LEVEL 219 

venient for reference. It should be noted from the figures in 
columns 4 and 5, which are for the ordinary range of press- 
ure employed in mining, that, though there is a difference of 
20 lbs. between the two gauge pressures, yet the outputs at dif- 
ferent altitudes vary only by a few thousandths and may 
often be neglected.* Wide differences, however, occur in the 
columns of mean effective pressures and horse-powers. 

Table XIII 





Barometric 




1 






Cubic Feet Pis- 


Cubic Feet 


& 




Relative Out- 
put for Gauge 


M. E. P. for 


ton Displace- 
ment per I. H.P. 


Comp 
Air per 


ressed 




D 


l h. p. 


1 


</)^ 


rS rt 


Pressure. 






for Gauge Pres- 


for Gauge 


1 


m 










sure. 


Pressure. 


< 


70 lbs. 


go lbs. 


70 lbs. 


90 lbs. 


70 lbs. 


90 lbs. 


70 lbs. 


90 lbs. 


1 


3 


3 


4 


6 


6 


7 


8 


9 


10 


11 





30.00 


14-75 


1. 000 


1. 000 


33'^ 


38.2 


6-93 


5-99 


1. 144 


.801 


1,000 


28.88 


14.20 


.967 


.966 


32-6 


37-6 


7-03 


6.09 


1. 123 


.787 


2,000 


27.80 


13.67 


•935 


-933 


32.1 


36.9 


7-15 


6.20 


1. 103 


-773 


3,000 


26.76 


13.16 


.904 


.900 


31-5 


36-3 


7.27 


6.31 


1.084 


-759 


4,000 


25.76 


12.67 


-«73 


.869 


31.0 


35-6 


7-39 


6.43 


1.065 


.746 


5,000 


24.79 


12.20 


-843 


.»39 


30-5 


35-0 


7-51 


6-55 


1.046 


-733 


6,000 


23.86 


11-73 


.813 


.809 


30.0 


34-3 


7-65 


6.67 


1.028 


.720 


7,000 


22.97 


11.30 


-7H5 


.780 


29.4 


33-7 


7.80 


6.79 


I. Oil 


.708 


8,000 


22.11 


10.87 


-75« 


-751 


28.9 


33-1 


7-94 


6.92 


-994 


-695 


9,000 


21.29 


10.46 


•731 


-723 


28.3 


32-5 


8.09 


7.06 


,976 


.683 


10,000 


20.49 


10.07 


•705 


.696 


27.8 


3i-« 


8.24 


7.20 


.959 


.670 


11,000 


19.72 


9.70 


.680 


.671 


27.4 


31.2 


H-39 


7-34 


.942 


.658 


12,000 


18.98 


9-34 


.656 


.647 


20.9 


30.6 


8.54 


7-49 


-925 


.646 


13,000 


18.27 


8.98 


.632 


.623 


26.3 


30.0 


8.71 


7-64 


.908 


-635 


14,000 


17-59 


8.65 


.608 


.600 


25.8 


29.4 


8.88 


7.80 


.891 


.624 


15,000 


16.93 


8.32 


•5S5 


•576 


25-3 


28.8 


9.06 


7-96 


.875 


-613 



Owing to the increase of piston displacement per indicated 
horse-power, as shown in columns eight and nine of the table, 
some builders make the air cylinders of compressors for mountain 
work of larger diameter for the same size of steam cylinder than 
those for sea -level service. As against the losses of the air end of 
the compressor at high altitudes, there is some gain in mean effec- 
tive pressure of the steam cylinders, because the exhaust takes 

* Attention may be called to the fact that for this reason, in compressor-builders' 
catalogues, no account is taken of the gauge pressures in tables of compressor 
capacities at altitudes. 



2 20 COMPRESSED AIR PLANT 

place against lower atmospheric pressure. The same is true in 
part of the air exhaust of machines using the compressed air. 
But the resultant of these gains is small and cannot be given much 
weight in offsetting the losses. A large deduction, for example, 
would have to be made for the lower calorific power of a given 
fuel at high altitudes. 

The relation between compressor output and barometric 
pressure may be expressed simply in another way. Take the case 
of two compressors of the same size, one operating under an 
atmospheric pressure of, say, 14 lbs. and the other at 10 lbs. (cor- 
responding approximately to an altitude of 10,000 ft.). If the 
first compressor is producing 6 compressions, the final absolute 
pressure will be 14 X 6 = 84 lbs. or about 70 lbs. gauge pressure. 
To produce the same gauge pressure, the other compressor must 
work to an absolute pressure of 70 + 10 = 80 lbs., the number of 
compressions corresponding to which is f-g- = 8. From each cubic 
foot of free air the first compressor will produce -|- of a cu. ft, 
of compressed air, and the second compressor, ^ cu. ft. Hence, 
the ratio of the respective outputs of the two compressors will 
be|- -^ -g- =f or 0.750. As compared with this, the ratio of the 
respective barometric pressures isi|- = o.7i4. 

Mechanically Controlled Inlet Valves for High Altitudes. It 
is often stated that compressors whose inlet valves are under some 
mechanical control are of special advantage for work at altitudes 
above sea-level. While there is a measure of truth in this, the 
possible saving is necessarily small, except at considerable eleva- 
tions. The question presents itself as follows. If the valve re- 
sistance be diminished by introducing mechanical control, so that, 
under normal conditions at sea-level, the inlet air will begin to 
enter the cylinder a little earlier in the stroke, the volumetric 
capacity of the compressor is thereby increased. The loss of 
capacity due to resistance of the valve springs, etc., which has 
been assumed to be 0.75 lb., for ordinary poppet valves, is a 
constant, and therefore becomes proportionately of greater and 
greater consequence as the altitude increases, because its ratio to 
the diminishing atmospheric pressure goes on increasing. The 



AIR COMPRESSION AT ALTITUDES ABOVE SEA-LEVEL 221 

percentage of saving obtained by eliminating the spring resistance, 
though small at or near sea-level, therefore becomes a matter 
of importance at great elevations ; and the inlet valve which pre- 
sents the smallest resistance to the entrance of the air into the 
cylinder will be the most economical for service in high moun- 
tain regions. 

Stage-Compression at High Altitudes. According to the state- 
ment already made, the greater the altitude above sea-level 
the smaller will be the ratio between the final pressure at delivery 
and the atmospheric pressure; that is, the ratio of compression. 
In Chapter V the effect of clearance in the air cylinder was dis- 
cussed, and it is evident that the percentage loss from this cause 
increases with the altitude because the piston must advance 
farther before the clearance air has been re-expanded to a press- 
ure below the diminished atmospheric pressure. Even if it be 
questioned whether it is worth while at sea-level to adopt stage 
compression for the ordinary pressures used in mining and 
tunnelling, the case is materially altered at high altitudes. 
For example, if it be desired to produce a gauge pressure of 
75 lbs. at 5,000 ft. elevation, corresponding to an atmos- 
pheric pressure of about 12.2 lbs., 7.15 compressions 
are necessary. At sea-level this number of compressions 
would give a gauge pressure of (14.7 X 7.15) — 14.7 = 90.4 
lbs. So far as losses due to piston clearance are concerned, 
therefore, it is as reasonable to employ stage-compression 
for 75 lbs., at 5,000 ft. elevation, as for 90 lbs. at sea-level. 
In a compound compressor, too, it must be remembered that there 
is practically but one clearance space: that in the intake cylinder. 
The value of the intercooler also increases with the altitude 
because, in beginning compression at an initial pressure below the 
normal, the greater total range of pressure through which the air 
must be carried involves the production of more heat. This 
additional heat must be effectually dealt with by the cooling ar- 
rangements, if loss from this cause is to be avoided. 

Considered from both the economic and thermodynamic stand- 
points, there can be no question as to the value of stage compres- 



222 COMPRESSED AIR PLANT 

sion for high altitudes. There is not only a decrease in output 
and an increase in the cost of production of the air, due to the 
added power required; but, as a result of these conditions, the 
compressor itself must be larger for a given output, and therefore 
its first cost will be greater than that of a compressor of the same 
capacity, working under normal atmospheric pressure. Hence, 
by introducing stage compression a larger percentage of saving 
is possible at high altitudes than at sea -level. 



CHAPTER XIV 

EXPLOSIONS IN COMPRESSORS AND RECEIVERS 

Explosions in air compressors and receivers occur with suf- 
ficient frequency to demand careful attention. Though they arc 
unquestionably attributable to ignition of volatile constituents 
of the lubricating oil, the immediate causes leading to this com- 
bustion are not always, nor altogether, clear. It is found, however, 
that explosions occur only in dry compressors, and some light 
may be thrown upon the subject by considering the conditions 
affecting the use of lubricant in these machines. In Chapter V 
attention was called to the fact that, if the cylinder temperature of 
a dry compressor be allowed to rise too high, not only does proper 
lubrication become difficult, but the oil itself may be decomposed 
by the heat. It is probable that ignition unattended by actual 
explosion is of frequent occurrence. Instances are on record 
where the discharge pipe near the compressor has become red- 
hot, and the ignition even extended into the receiver without pro- 
ducing a destructive explosion. Examination of the discharge- 
valve chests and passages, and the pipe leading from compressor 
to receiver, often reveals the presence of a black, sooty residue 
originating from decomposition of the lubricant. The volatile 
constituents of the oil thus liberated, on passing with the com- 
pressed air into the receiver, would make a mixture of air and gas 
capable of producing an explosion. The extreme violence often 
noted in such explosions is probably due in part to the high air 
pressure existing in the valve passages, discharge pipe, and re- 
ceiver. In high pressure air, combustion is always more active 
than in air at atmospheric pressure. 

A number of the recorded air-compressor explosions have oc- 

223 



224 COMPRESSED AIR PLANT 

curred at collieries, and the possible effects of the presence of 
coal dust in the intake air of the compressor have been carefully 
considered. A deposit of such dust in the valve passages, together 
with the sooty residue from decomposition of the oil, might as a 
result of oxidation produce a condition very favorable to an ex- 
plosion. It has been suggested that, in these circumstances, a 
spark caused by the friction of the compressor piston, if work- 
ing dry, might bring about an explosion; or, by the continual 
passage of air at a high temperature over the carbonaceous de- 
posit, spontaneous combustion might result, and ignite the in- 
flammable mixture of oil-vapor and air.* However, there are a 
sufficient number of cases where explosions have taken place at 
mines and works other than collieries to prove that such explo- 
sions are not necessarily dependent upon the presence of coal-dust 
in the intake air of the compressor. When the compressor is 
improperly situated in a room close to the boilers and coal-bins 
some coal dust might be present in the air; but though possibly 
assisting in the explosion, the quantity could hardly be large 
enough to produce by itself the observed results. 

The true cause of these explosions is undoubtedly to be found 
in the working conditions prevailing in the compressor cylinder. 
In a single-stage dry compressor an excessively high tempera- 
ture is often reached, because of improper design of the air 
cylinder, or by running too fast (as when the compressor is too 
small for its work), or by attempting to produce too high a 
pressure. The temperature of the discharge air from a single- 
stage compressor is found by the formula already given in 
Chapter X: 

F 

P 

in which : T and P are, respectively, the absolute initial temper- 
ature and pressure of the intake air; T' and P', the absolute final 
temperature and pressure; and n, the constant, 1.41. Under 
normal conditions near sea-level, say, when the temperature of the 
atmosphere is 70° F., P = 14 lbs., and the gauge pressure at dis- 

* T. G. Lees, Trans. Federated Inst. Mining Engineers, Vol. XIV, p. 568. 



(1)-- 



EXPLOSIONS IN COMPRESSORS AND RECEIVERS 225 

charge, 80 lbs., the final temperature is found by making the 
respective substitutions, 

whence T' = 70 + 459 ° [ -—) =917° F. absolute, 

or 458° F. by the thermometer. 

As calculated by this formula, the compression is supposed 
to be purely adiabatic, no account being taken of loss of heat by 
radiation or of any effect that may be produced by the water- 
jackets. As a matter of fact, but little heat can be abstracted by 
the jackets of a single-stage compressor. Air is a poor conductor, 
and the volume in the cylinder is not long enough under the in- 
fluence of the jackets to be much affected by them. In a com- 
pressor of this type the chief ofhce of the jackets is to keep down 
the temperature of the cylinder walls and prevent the lubricating 
oil from being carbonized. It is probable, therefore, that in a 
single-stage dry compressor, even if well designed and in good 
order, the actual temperature of the air at discharge will generally 
range from, say, 375° to 425° F., and may often go higher — 
a statement sufficiently supported by recorded observations. 

In consideration of what precedes it is evident that the quality 
'of the lubricating oil used in the air cylinder, and especially its 
flashing- and ignition-points, are matters of importance.* The 
flashing-point of ordinary cylinder oil may be taken as from 330° 
to 425° F. "An average of determinations on 40 samples of 
heavy oils having an average flash-point of 360° F., gave average 
burning-point of 398° F. High flash test cylinder oils, from 500° 
to 560° F., gave burning-points of 600° to 630° F." t Common 
lubricating oils flash at about 250° F., and kerosene, sometimes 
carelessly used by compressor engineers for cleaning discharge 
valves, at 150° F. or below. In the case of one explosion the 
flash-point of the cyhnderoil used was found to be only 295° F. J 

* The flashing-point of oil is the lowest temperature at which it gives off com- 
bustible vapors in sufficient quantity to be ignited by contact with flame. The 
ignition-point is the temperature to which the vapors must be raised in order to 
continue to burn. 

t Alex. M. Gow, Engineering News, March 2d, 1905, p. 221. 

f John Morison, Trans. North oj England Inst. Min. Engs., Vol. XXXVIII, p. 6. 



226 COMPRESSED AIR PLANT 

It would appear, from a comparison of these temperatures, that 
an explosion in a compressor cylinder, directly traceable to de- 
composition of the lubricant, would be possible under normal 
conditions only when inferior, light mineral oils are employed. 

But compressors are not always in good order, nor the work- 
ing conditions always normal in other respects. Aside from the 
dangers arising from the use of low-grade lubricant, it is more 
than probable that one of the commonest causes of explosion is 
air-cylinder leakage, either of the delivery valves or past the pis- 
ton. The effects of leakage may be illustrated by citing a case or 
two. 

In 1897 an explosion took place in one of the receivers of the 
compressor at the Clifton Colliery, England.* It attracted much 
attention, and is so instructive that many of the details are given 
here. The air from the compressor passed to a series of 3 
receivers of large size, the first being 7 ft. diameter by 40 ft. 
long. While running apparently under normal conditions the 
safety-valves of the receivers suddenly began blowing off with a 
deafening roar. Flames several feet high issued at great press- 
ure from the safety-valves, and sparks were blown out at the 
joints of the 8-in. pipe leading from the compressor to the 
first receiver. The air main near this receiver was nearly red- 
hot. That the receivers did not burst was thought to be due 
to the relief afforded by the 4 safety-valves — 2 on the first 
receiver and i on each of the others — and to the fact that the 
underground engines driven by compressed air continued running 
for some minutes after the compressor was stopped. On ex- 
amining the first receiver, after it had cooled, it was found that, 
just below the point at which the air entered from the compressor, 
a mass of black carbonaceous matter had been deposited, from 
I J to 2 ins. thick and 6 sq. ft. in area. On analysis this 
showed: volatile matter, 55.8 per cent., fixed carbon, 37.3 per 
cent., and ash, 6.9 per cent. The material was charred and 
had the appearance of hard vulcanite. A thin coating was 
noticed on the sides of the receiver (though only near the inlet 

*T. G. Lees, Trans. Federated Inst. Mining Engineers, Vol. XIV, pp. 555-559. 



EXPLOSIONS IN COMPRESSORS AND RECEIVERS - 227 

pipe) and also in the air pipe itself. The other two receivers were 
free from deposit. A coating of carbonaceous matter, to a thick- 
ness of one-quarter inch was found on the discharge valves and 
passages. The cylinder and piston surfaces were not dry and, 
though they showed signs of excessive heat, were uninjured. 

The gauge pressure was usually 60 lbs., which, with adia- 
batic compression, would correspond theoretically to a final 
temperature of 405° F., the temperature of the intake air from 
the engine-house being 80°. The lubricating oil used was 
guaranteed to have a flash-point of 554°, and ignition-point of 
606° F. As the cylinders were water-jacketed, the actual final 
temperature should not, in regular working, reach the above- 
named temperatures; in fact, readings previously taken from a 
thermometer in the outlet pipe showed that it usually registered 
about 350° F. It is significant, however, that on one occasion the 
mercury rose above 500°, and the thermometer tube burst. The 
temperature at the time of the explosion therefore was not known. 
Afterward a pyrometer was fixed on the outlet pipe as near as. 
possible to the discharge valves, and the temperature was found 
to range generally from 400° to 420° F., varying with the speed of 
the engine and the air pressure produced. Even with these tem- 
peratures, high as they are, it would seem impossible that ignition 
of the lubricating oil could take place. It is evident that an un- 
usual increase of temperature in the air cylinders must be ac- 
counted for. 

In commenting on this accident, Mr. W. L. Saunders makes 
the following interesting remarks on explosions in compressors 
and receivers : 

'' There must be an increase of temperature, or ignition would 
not take place. This increase of temperature may result either 
from an increase of pressure, which is not recorded on the gauge, or 
there may be an increase of temperature without a corresponding 
increase of pressure. Take the first instance, and it is not dif- 
ficult to understand that a compressor might deposit carbon from 
the oil in the discharge passages or discharge pipes, which in the 
course of time will accumulate and constrict the passages so that 



228 COMPRESSED AIR PLANT 

they do not freely pass the volume of air delivered by the com- 
pressor. Hence, a momentary increase of pressure might exist in 
the cylinder heads, or in the discharge pipe which leads from the 
cylinder to the receiver, which would surely carry with it an in- 
crease of temperature possibly exceeding the ignition-point of 
the oil. A badly designed compressor with inefficient discharge 
passages might produce this trouble. Too small a discharge 
pipe or too many angles in discharge pipes might also tend to 
produce explosions. But ignition is known to have occurred in 
a well-designed system, and other causes must be sought. We 
think many cases may be traced to an increase of temperature 
without increase of pressure; this increase of temperature can be 
excessive only when the temperature of the incoming air is ex- 
cessive. A hot engine-room from which air is drawn into the 
cylinder is a bad condition. Ignition is known to have taken 
place, however, when the temperature of the incoming air was 
normal, when the discharge passages and pipes were free and of 
ample area, so that some other cause must still be looked for. 
The only possible explanation is that the temperature of the in- 
take air is made excessive by the sticking of one or more of the 
discharge valves, thus letting some of the hot compressed air back 
into the cylinder to influence the temperature before compression. 
... It is not difficult to understand a leaky discharge valve 
letting enough hot compressed air back into the cylinder to in- 
crease the initial temperature to 200 or 300°. If so, and the 
air is being compressed to 73.5 lbs. gauge pressure we have, 
say, 300° temperature in the free air before compression, and as 
the increase is 354.5°, the resulting temperature might be 654.5°, 
As a remedy we would suggest more care in selecting the best com- 
pressor, and in frequent cleaning of the discharge valves and 
passages. The best compressors are built so that the discharge 
valves may be readily removed. These valves should be cleaned 
once a week by the engineer, who should see that they fit properly. 
It is impossible to get good lubricating oil that is free from carbon, 
hence there will always be more or less carbon deposited on the 
discharge valves, but this must not be allowed to accumulate. 



EXPLOSIONS IN COMPRESSORS AND RECEIVERS 229 

Intercoolers between air cylinders and aftercoolers between 
final cylinder and receiver are also recommended. One of these 
coolers located in the discharge pipe will absolutely prevent the 
passage of flame, and will insure the protection of the mine 
against fire even though there be ignition at or near the air 
cylinder." * 

During the construction of the New York Aqueduct a fire oc- 
curred in a compressor receiver at one of the shafts. The air 
pressure was eighty to ninety pounds, and the horizontal receiver, 
set outside of the engine-house, was exposed to the hot sun. Part 
of the discharge pipe leading to the receiver had become red-hot. 
On stopping the compressor and cooling down the receiver, the en- 
tire inner surface of the latter was found to be coated with carbo- 
naceous matter at least one-eighth inch thick. Further investiga- 
tion brought out the fact that the poppet discharge valves had 
sometimes occasioned trouble by sticking, and the engineer had 
been in the habit of using a squirt-can of kerosene to cut the 
gummy material clogging them. As the kerosene had a low flash- 
point, it was quickly vaporized, and when the cylinder tem- 
perature reached a sufficiently high point the explosion took place. 

In this case, as in that previously cited, the trouble seems^ to 
have been caused by leakage of the delivery valves (possibly past 
the piston also), thereby raising the cylinder temperature to an 
abnormal degree. It may be added that the use of kerosene 
for cleaning gummy discharge valves is a dangerous practice, even 
when the compressor is slowed down while using it. 

The effect of leakage in the air cylinder may readily be under- 
stood from the following discussion of what takes place in the 
course of a single stroke, with the accompanying temperature 
changes. t At the beginning of the stroke, the air in the cylinder 
consists of: that which remained in the clearance spaces at the 
end of the previous stroke, that which has leaked in, and that which 
has been drawn in from the atmosphere. The clearance air, on 

* Compressed Air, July, 1897, pp. 258-259. 

t Abstracted from a paper by E. Hill, Trans. Amer. Inst, of Min. Engs, Vol. 
XXXIV, p. 950. 



230 COMPRESSED AIR PLANT 

re-expanding, falls from an absolute temperature of T' to T (see 
formula near the beginning of this chapter), and its effect may 
therefore be neglected. For well-designed compressors, the tem- 
perature of the air newly drawn into the cylinder may be taken as 
that of the outside atmosphere, /, though it is generally heated 
in some degree by contact with the hot inner surfaces of the 
cylinder. Finally, if L represent the volume of air leakage, then, 
since T is the absolute temperature of the entire mass of air 
occupying the cylinder at the beginning of the stroke : 

T = (i-L)/ + TL (i) 

If, in the expression previously given for the temperature of the 
discharge air, viz : 

/p'\o.29 

t'=t(p) 

the pressures be written in atmospheres; then, for compressors 
working at sea-level, P = i and : 

T'= T P' -^, whence T = p^ (2) 

Placing the values of T, in equations (i) and (2) equal to each 
other and transposing : 

,_ /(P°-^-LF°-^) 
I - L P'°-29 

Applying this formula to a single-stage compressor, working 
to say, 7 atmospheres, or about 88 lbs. gauge, the atmos- 
pheric air being at 62° F., the discharge temperatures for differ- 
ent percentages of leakage will be as shown in Table XIV. The 
temperatures for an altitude of 4,000 ft. are also given for pur- 
poses of comparison. The leakages are expressed as percentages 
of cylinder capacity. 

These possible temperatures are fully sufficient to produce an 
evolution of gas, or even decomposition of the cylinder oil, causing 
it to burn; which would be followed by an increased liberation of 
volatile matter and the probability of explosion. 

It must be borne in mind, as pointed out by Mr. E. Hill, 
that leakage from an imperfectly fitting discharge valve is a con- 
stant in any given case, while the volume of intake air varies with 



EXPLOSIONS IN COMPRESSORS AND RECEIVERS 



231 



the speed of the compressor. Thus, a leak of 2 per cent, of the 
intake volume, at, say, 125 revolutions per minute, becomes 10 
per cent, if the compressor be slowed down to 25 revolutions. This 
agrees with experience, violent explosions being known to have oc- 
curred while the compressor was running slowly. " The oil-feed 
was probably adjusted to the maximum speed and hence was 
excessive for the slow speed. A larger proportional leak — a 
liberal quantity of oil — and the result is easily comprehended/* 



Table XIV 





Temperature of Discharge. Degrees Fahrenheit. 


Leakage. Per Cent. 






At Sea-Level. 


At 4,000 Feet Elevation. 





459 


496 


I 


466 


504 


2 


475 


513 


4 


489 


530 


6 


506 


549 


8 


524 


570 


10 


544 


593 


12 


566 


618 


14 


589 


646 


16 


615 


675 



The effect of leakage of the discharge valves, moreover, is cumu- 
lative, for each rise in initial temperature thereby produced 
causes a greater rise in terminal pressure; and the leakage con- 
tinuing, a very few strokes would suffice to ignite the best cylinder 
oil. In some circumstances, even a single stroke of the piston 
may cause ignition, if not explosion. 

The importance of minimizing piston and discharge-valve 
leakage is evident. One of the surest means of avoiding danger 
of high cylinder or receiver temperatures is the adoption of stage 
compression. There are two reasons for this: (i) the air is partly 
cooled between the stages, so that the maximum temperature is 
always less than in single-stage machines, and (2) the leakage is 
likely to be less because there is a smaller difference between the 
pressures on the two sides of the piston, as well as between the 
internal and external pressures on the discharge valves. 



232 



COMPRESSED AIR PLANT 



A case of explosion, in which the influence of cylinder leakage 
is not clearly apparent, occurred some years ago in the air pipe of 
a large plant in Butte, Mont. Two duplex compressors, with 
air cylinders respectively of 32 J X 60 ins. and 24 J X 48 ins., 
and running at 50 revolutions per minute, were forcing air at 80 
lbs. pressure through a single 8-in. pipe. As somewhat over 
1,200 cu. ft. of compressed air per minute were being pro- 
duced, the velocity of flow would be nearly 3,500 ft. per min- 
ute, or 58 ft. per second. It had been noticed several times 
that a portion of the discharge pipe close to the compressor be- 
came red-hot. As no explosion took place in the compressor 
cylinders, but in the pipe only, it is probable that the oil accum- 
ulated in the pipe was vaporized and ignited. In the pipe between 
compressors and receivers there were several sharp bends, which 
increased the friction due to the rapid flow of the air. The re- 
ceivers were always extremely hot. On one occasion the shaft 
timbering, forty or fifty feet below the shaft mouth, took fire from 
the hot air pipe. The above gives point to the fact that, while the 
primary causes of explosion are to be found in the air cylinder, 
the disastrous effects are perhaps oftener observable in the dis- 
charge pipe or receiver. 

Foul or poisonous gases may result from ignition of the 
lubricant in compressors or receivers, not necessarily followed by 
actual explosion. In an article in the Trans, Amer. Inst, Min, 
Engs.j Vol. XXXIV, p. 158, an instance is noted of combustion in 
the air pipe and receiver. The compressed air was being used 
in an imperfectly ventilated upraise in the mine, 1,200 ft. from 
the compressor, and 2 men lost their lives, while 4 others barely 
escaped asphyxiation. 

Other more or less similar cases are familiar to most miners, 
where foul air from the exhaust of machine drills has been ob- 
served; sometimes merely disagreeable, though often actively 
deleterious. The use of poor cylinder ofl is frequently responsible 
for this, as its lighter constituents may begin to volatilize and burn 
at a perfectly normal working temperature. Even if not actually 
fried on the hot metal surfaces, a low-grade oil may yet undergo 



EXPLOSIONS IN COMPRESSORS AND RECEIVERS 233 

a slow combustion or oxidation, which will produce enough car- 
bon dioxide to raise materially the percentage of that poisonous 
gas in the confmed atmosphere of the working places of mines. 

The mode of using the lubricant for the air cylinders of com- 
pressors deserves some attention. Sight-feed lubricators, such 
as are commonly employed for steam cylinders, are best. On 
the Clifton Colliery compressor, mentioned above, ordinary 
oil-cups were used, holding about |- pint. They were filled 
4 times per day of 10 hours. With these oil-cups, if improp- 
erly adjusted, it would be possible for all the oil to be sucked 
into the cylinder within a few strokes after being filled. Such 
a result might be inferred, indeed, in this case, because of the 
largequantityof carbonaceous matter — oil, coal dust, etc. — found 
in and around the discharge valves and in the receiver. The feed- 
ing of the oil should be carefully regulated, and a smaller quantity 
used in an air cylinder than a steam cylinder of the same size — 
say, one-third as much. An excess of oil increases the tendency 
to gum the valves. For stage compressors of ordinary size, 
I drop of good cylinder oil every 4 to 5 minutes is sufficient. 

The periodical use of soap and water (soap-suds) is to be 
recommended for any compressor that cannot be shut down at 
short intervals for overhauling. It is fed into the air cylinder 
through an oil-cup, say during one day per week. Or it may be 
forced in by an oil-pump, with which the air cylinder should be 
provided. Soap and water is a poor lubricant in itself, and must 
be used more freely than oil, but it is effectual in cleansing the 
cylinder, valves, and ports from any carbonaceous or gummy 
matter that may have been deposited. If the compressor is to be 
stopped, as at the end of a shift, care must be taken to discontinue 
the feeding of soap and water some time before shutting down, 
and resume the oil-feed. This is necessary to avoid the formation 
of rust. Every compressor should be overhauled from time to 
time, and thorough cleaning should extend to all parts, especially 
around the valves and passages, capable of furnishing a lodgment 
for oil or partly oxidized carbonaceous material. 

Precautions for Preventing Explosions. These may be sum- 



234 COMPRESSED AIR PLANT 

marized as follows: (i) Always enclose the inlet valves in a cold- 
air box, connecting with the outside air, so as to avoid taking the 
air from the hot engine-room. This not only conduces to econ- 
omy in working, but by keeping down the final temperature tends 
to prevent decomposition of the oil. (2) The largest possible area 
of cylinder surface should be water-jacketed, including the cylin- 
der heads. A liberal supply of the coldest water obtainable 
should be used for the jackets. The advantages in this respect 
derived from employing stage compression, with large inter- and 
aftercoolers, are undoubted. (3) Use only the best cylinder oil, 
with high flash- and ignition-points and in as small quantity as is 
consistent with proper lubrication. Care should always be taken 
to keep the valves clean. In the design of the compressor there 
should be no recesses or pockets, around the valves or passages, 
where oil could accumulate. (4) So arrange the air intake that 
coal dust will not be drawn into the cylinder with the inlet air. 
(5) It is well to place a thermometer in the discharge pipe, 
close to the cylinder, so that the engineer will be able to note the 
temperature from time to time and stop or slow down the com- 
pressor if the temperature of the discharge air rises too high. 



CHAPTER XV 

AIR COMPRESSION BY THE DIRECT ACTION OF 
FALLING WATER 

In view of the economic importance of keeping down the tem- 
perature of the air during compression, it is evident that an ad- 
vantage would be derived from a closer and more intimate con- 
tact between the air under compression and the cooling water 
than is possible with the external water-jackets of dry compress- 
ors. From a thermodynamic standpoint it cannot be questioned 
that the wet compressor is more efficient than the dry. As has 
been shown in the latter part of Chapter V, it is mainly the me- 
chanical difficulties resulting from the use of injected water in 
the air cylinder that operate to the disadvantage of the wet system 
of compression, and that have caused its almost complete aban- 
donment. 

Since 1896 several large plants have been successfully in- 
stalled in which air is compressed by the direct action of falling 
water and without the use of piston, valves or other moving parts. 
The simple and familiar principle involved has aroused much 
interest in this method of air compression. When air in small 
bubbles is intimately mixed with water, the water breaks into 
foam, through which the air bubbles tend to rise and escape. But 
if the mixed air and water be drawn downward by a strong falling 
current, suitably confined, as in a vertical pipe, the air is com- 
pressed. And if, after reaching the depth and head of water col- 
umn necessary to produce the degree of compression desired, the 
direction of flow be changed to the horizontal and the velocity 
diminished, the air bubbles will rise. They may then be collected 
in a suitable chamber, in which the air pressure corresponds to 
the head of water and from which the air is drawn off as required. 

235 



236 COMPRESSED AIR PLANT 

As the air bubbles are minute and thoroughly disseminated 
through the water during its descent, the total cooling surface 
presented is very large and complete isothermal compression re- 
sults. It should be observed also that the compressed air is 
very dry. While undergoing reduction in volume, the percen- 
tage of moisture in a given globule of air increases until the point 
of saturation is reached, but any further compression causes 
deposition of part of the moisture. Moreover, since the air is 
kept constantly cool during compression, its moisture-carrying 
capacity is smaller than if compressed adiabatically, as in an 
ordinary compressor cylinder. 

Although such an apparatus embodies no new principle, it was 
first constructed on a working scale and successfully tested, about 
1878, by J. P. Frizell, of Boston, Mass.* Aided by this prece- 
dent, a more effective and practical method of breaking up the 
water and impregnating it with air in a state of fine division, was 
afterward devised by Charles H. Taylor, of Montreal, Canada. 
In 1896 the Taylor Hydraulic Air Compressing Co., of Mon- 
treal, erected a plant embodying the system for the Dominion 
Cotton Mills, Magog, Province of Quebec. f This plant has long 
been in successful operation, and where the conditions permit its 
introduction the system may be advantageously employed for 
mining service also. 

For the Magog Mills a 128-ft. shaft was sunk to give the 
desired head and pressure (Fig. 117). In it was erected a large 
vertical compressing pipe, a, 3 ft. 8 J in. diameter, the lower part 
gradually increasing to 4 ft. 8 in., and made of T^-in. steel plate. 
This pipe passes through the bottom of an iron receiving cham- 
ber, 6, at the surface, to which water is conducted from a dam 
or reservoir. The chamber, h,h 12 ft. diameter by 12 ft. high. 
Water flows into and fills the pipe, which extends nearly to the 

* For a record of these tests see Proceedings of the Institution of Civil Engineers, 
London, Vol. LXIII, p. 347. 

t The following description is based on an article in the Canadian Engineer, 
March, 1897, and information furnished to the author by the builders. See also 
Eng. and Mining Jour., Dec. 26th, 1896, p. 606, and Railway and Engineering 
Review, Sept. 17th, 1898, p. 513. 



AIR COMPRESSION BY ACTION OF FALLING WATER 237 




Fig. 1 1 7. — Taylor Hydraulic Air Compressor. 



22>^ 



COMPRESSED AIR PLANT 



bottom of the shaft. By means of an arrangement of small feed 
pipes described below, air is drawn with the water into the top of 
the main vertical pipe and is compressed while being carried 
down the shaft. The compressed air collects in a separating 



Plan of Spider 

of Cylindrical 

Head Piece 




Fig. II 



chamber, c, at the bottom of the shaft, while the water is returned 
up the shaft to a tailrace near the top. The difference of water 
level between intake and tailrace is about 22 ft., which produces 
the requisite speed of flow of the mass of water. Into the top 



AIR COMPRESSION BY ACTION OF FALLING WATER 239 

of the vertical pipe, <2, is inserted a telescoping section of pipe, 
(/, to the upper end of which is riveted a bell-mouth, e. 
Above the latter is a cylindrical headpiece, /, 4 ft. 8 ins. diameter 
(Fig. 118), terminating below in an inverted conoid, ^, projecting 
into the bell-mouth. These two parts are connected by lugs 
and bolts in such way as to leave an annular opening between 
them, through which the water enters the vertical pipe. Around 
the headpiece is set a series of thirty 2 -in. pipes, h^ h, 4 ft. long, 
open at the top and closed at the bottom. Into each of these 
pipes, near their lower ends, are screwed 32 short horizontal |-in. 
pipes, i, z, all directed into the annular opening at the bell-mouth 
and toward the axis of the main pipe. As the entering water passes 
among the small pipes a tendency to vacuum is created in them, 
so that the atmospheric pressure drives the air through them into 
the water in the form of small bubbles. These are carried with 
the water down the main pipe, and on their way are compressed. 
Near the bottom of the shaft the vertical compressing pipe 
enters the large circular '^ separating " chamber, c, 17 ft. diam- 
eter and 12 ft. high, open below and supported upon legs which 
raise it 16 ins. above the shaft bottom. Within the tank and di- 
rectly under the pipe is the "disperser," /, a conoidal casting like 
the one in the headpiece. Plates, ^, are added around the periph- 
ery of the disperser to give it an outside diameter of 12 ft. 
Below is an inverted conical apron, /, 5 ft. wide, riveted to the 
interior of the separating tank. When the water, charged 
with air bubbles, reaches the disperser it is directed outward 
toward the circumference; is then deflected by the apron tow- 
ard the center under the disperser, and finally escapes through 
the open bottom of the separating tank into the return column. 
During this process of travel the compressed air separates from 
the water, most of it collecting in the upper part of the air cham- 
ber. A portion of the air is not liberated until the water reaches 
the lower part of the tank, under the apron. This residuum 
collects in the annular space and joins the main body of air 
through the pipe, m. The compressed air collecting in the top of 
the air chamber is kept under pressure by the weight of the re- 



240 



COMPRESSED AIR PLANT 



turn water column in the shaft, and is drawn off through the ver- 
tical air main, alongside of the water column a. As the small 
air bubbles are constantly surrounded by cold water, it is evi- 
dent that by this system perfect isothermal compression is at- 
tained, with its corresponding advantages in minimizing the 
amount of moisture carried off in the air. This has been shown 
by tests. 

With a total depth of shaft of 128 ft., in this installation, an 
air pressure of 52 lbs. per sq. in. is produced. The efficiency of 
this plant is shown by the following table* to be from 50.1 per 
cent, to 62.4 per cent., according to the quantity of water used: 

Table XV 



No. 

of 

Test. 



Quantity 
of Water 

Dis- 
charged, 
in Cubic 
Feet oer 
Minute. 



Available 


Available 


Head in 


Hor«'^- 


Feet. 


Power. 


21.4 


247.7 


21.9 


228.0 


22.3 


168.9 


21. 1 


305-9 


21.7 


260.0 


21.2 


299.8 



Quantity of Air 

Delivered, in 

Cubic Feet per 

Minute at 

Atmospheric 

Pressure. 



Pressure 
of Air, 
Pounds 

per 
Square 
Inch. 



Actual 
Horse- 
Power of 

Com- 
pressor. 



Efficiency 
of Com- 
pressor, per 
Cent. 



6122 

5504 
4005 
7662 
6312 
7494 



1377 
1363 
1095 
1616 
1506 
1560 



52 

52 
52 
52 
52 
52 



132-5 
131. o 

105-3 
155-4 
144-8 
150.2 



53-5 
57-5 
62.4 
50.8 
55-7 
50-1 



Temperatures during tests: external air 75° to 83°; water 75.2° to 80°; 
compressed air 75.2° to 80°. 

The parts were not correctly proportioned in this first instal- 
lation, and there is no doubt that the efficiency could be consider- 
ably increased by using a relatively larger air chamber at the 
bottom of the shaft, to prevent air from going to waste. As shown 
by the table, the efficiency is increased by diminishing the volume 
of inlet water, upon which depends the quantity of air carried 
down and compressed. 

In building a plant to produce higher air pressure the motive 
head, or difference in level between the surfaces of water at inlet 
and tailrace, would be increased. The theory is as follows : The 

* Tests made by Prof. C. H. McLeod, of McGill University, August, 1896. Pub- 
lished in Eng. and Min. Journal, December 26th, i8g6, d. 606. 



AIR COMPRESSION BY ACTION OF FALLING WATER 24I 

combined specific gravity of the mixture of air and water in the 
vertical compressing pipe is less than that of the water in the re- 
turn column. That is, the weight of water in the compressing 
pipe is less per foot than in the return column. Therefore, the 
head required, to overcome friction and to produce flow, must be 
greater than if the apparatus were merely an inverted siphon, and 
as the difference in weight increases with depth (and air press- 
ure produced) the motive head must be correspondingly increased. 

In 1 898-1 900 a plant on the Taylor system was built for the 
Kootenay Air Supply Co., Ainsworth, British Columbia. The 
topographical conditions are such that a high head of water is 
obtained w^ithout sinking a deep shaft. From a small dam the 
water is carried in a wooden-stave pipe, 5 ft. in diameter and 
1,354 ft. long. The pipe finally passes over a short, but high 
trestle, built against the side of a steep gorge, to the receiving 
tank. The latter, 17 ft. diameter by 20 ft. high, is placed on a 
wooden tower, no ft. high (Fig. 119). From the bottom of the 
tank the pressure pipe, 33 ins. diameter, descends vertically inside 
the tower to the ground level and then down a shaft 105 ft. 
deep.* After compressing the air the water returns up the 
shaft to the tailrace at the creek level. As shown in Fig. 120, 
the details of the receiving chamber at the bottom of the shaft 
differ from those of the Magog plant. 

The effective compressing head is 107 ft., while the total 
height of the pressure pipe is over 200 ft. This produces a 
high velocity of flow and a correspondingly large delivery of 
compressed air. The main pipe line, 9 inches diameter, is 2 
miles long, discharging from 4,200 to 4,600 cu. ft. of free air 
per minute. Branch service pipes convey the air to neighboring 
mines, where it is used for rock-drills and other mining machinery. 
On the basis of 600 horse-power, represented by the volume and 
pressure of the air compressed, the cost of the entire plant, includ- 
ing pipe lines, was about $100 per horse-power. This would be 
somewhat increased by allowances for transmission and other 
losses. 

* Canadian Electrical News, September, 1898, p. 176. 



242 



COMPRESSED AIR PLANT 



Another large plant was completed in 1906 at the Victo- 
ria Copper Mine, near Rockland, Ontonagon Co., Michigan. 
Though the same general design was adopted for the intake head 





END ELEVATION 



SIDE ELEVATION 




PLAN 



Fig. 120. — Hydraulic Air-Compressor at Kootenay. 






j\ 


i 


SIDE ELEW 


SECTION ^ 


HROUGH A-a 

11 


=,^ 






■■' ^1| 


Itol 


i 


'^' 


,11 


^^fc^ 



Pig. 119, — Hydraulic Air Compressing Plant at Kootenay. 



AIR COMPRESSION BY ACTION OF FALLING WATER 243 

and its appurtenances, the local conditions led to a novel mode of 
installation. The water is conducted from a dam on the On- 
tonagon River through a 4,700-ft. canal, furnishing a head at 
the terminal forebay of 72 ft. above the river-level. Three 
independent units are built side by side in a vertical shaft 343 ft. 
deep. The subdivision of the air, as admitted at the intake head, 
is carried farther than in either of the plants described above, 
there being no less than 1,800 |-inch horizontal feed pipes, in- 
serted in the series of vertical pipes encircling the inverted cone. 
The compressing pipes are 5 ft. in diameter, lined with con- 
crete, and the separating cones and dispersers, also of iron and 
concrete, are built at the bottom in a rock chamber excavated for 
the purpose. In this chamber, 281 ft. long and 18 ft. X 21 ft. 
average cross-section, the compressed air is trapped and thence 
drawn off for use through a 24-inch main. The compress- 
ing water, flowing down the intake pipes, stands normally at 
a level about 14J ft. below the roof of the chamber, thus leav- 
ing an air capacity of about 80,000 cu. ft. Connected with the 
end of the air chamber is an inclined shaft, 270 ft. in vertical 
depth, through which the water returns to the surface. The tail- 
race from the mouth of this shaft is 72 ft. below the level of the 
intake, this height measuring the motive head producing the 
flow of water. Thus the air in the underground chamber is 
under a pressure due to 270 ft. head of water, or 117 lbs. per sq, 
in. gauge. 

For regulating the operation of the compressor a pipe passes 
from the air chamber up the compressing shaft to the surface, 
where branches from it are led to the intake heads. The com- 
pressed air conveyed in this regulating pipe operates a device con- 
nected with each intake head, whereby the latter is automatically 
raised above the water-level in the receiving tanks whenever the 
air pressure exceeds the normal, thus stopping the flow of air 
through the feed pipes. A twelve-inch blow-off pipe is also pro- 
vided, passing from the water-level in the air chamber to the mouth 
of the inclined shaft carrying the return water column. If air to 
the full compressor capacity is drawn off, the water-level in the air 



544 



COMPRESSED AIR PLANT 



chamber rises as the air pressure falls, thus sealing the lower end 
of the blow-off pipe; then, when the consumption of air decreases 
the pressure in the chamber rises, depressing the water-level until 
the blow-off orifice is uncovered, when more air is blown off. 
Thus the working pressure is maintained within quite narrow 
limits. It may be added that the great size of the air chamber — 
which acts like the receiver of an ordinary air-compressor plant — 
gives it a large storage capacity. 

When all 3 compressing units are in operation, with a total 
capacity of from 34,000 to 36,000 cu. ft. of air per minute, about 
70,000 cu. ft. of free air per minute may be drawn off for a 
period of 18 minutes, without causing a drop in pressure of more 
than 5 lbs. For each unit, the output ranges from 9,000 to 12,000 
cu. ft. per minute, and the volume of water used, from 12,700 
to 14,800 cu. ft. A series of tests made on a single intake head 
in May, 1906, by Prof. F. W. Sperr, gave the following results: * 



Table XVI 
Air Measurements 



Square Feet. 


Velocity, Feet 
per Second. 


Cubic Feet per 
Minute. 


Absolut 

Free Air, 
Pounds. 


E Pressures 

Compressed 
Air, Pounds. 


Horse- Power. 


4 
4 
4 


44.09 
49-74 
38-50 


10,580 

11,930 
9,238 


14 
14 
14 


128 
128 
128 


1,430 
1,623 
1,248 


Water Measurements 


Flume Area. 


Velocity, Feet 
per Second. 


Cubic Feet per 
Minute. 


Head, Feet. 


Horse- Power. 


Efficiency, per 
Cent 


71-75 
67.03 
72.16 


3-033 
3.684 
2.936 


13,057 
14,820 
12,710 


70-5 
70.0 
70.6 


1,741 
1,961 
1,700 


82.17 
82.27 
73-50 



The air is used at the Victoria Mine for general power pur- 
poses at the mine and mill, including a 500-horse-power hoisting 

* For further details see article by D. E. Woodbridge, Engineering and Mining 
Journal, Jan. 19th, 1907, p. 125. Also, A. H. Rose, Mme5 and Minerals, March, 
1907, P- 346. 



AIR COMPRESSION BY ACTION OF FALLING WATER 245 

engine, designed for a depth of 4,000 ft., 7 pumps, and many other 
engines. The cost per horse-power is only about $2.25 per 
year, including all the operating expenses. It is expected that 
over 4,000 horse-power will be developed when all 3 compressing 
units are in operation. The present stage of the development of 
the mine requires the use of but i unit.* 

The compression of air by direct action of falling water, 
according to the Taylor system, has been adopted in several other 
recent installations : two in Germany and a very large plant for 
general power purposes, on the Shetucket River, near Norwich, 
Conn.t It is probable that the appHcation of the system will be 
extended in regions where large water powers can be developed. 
Its first cost is not excessive, while the maintenance and running 
expenses are extremely low, as compared with those of the usual 
forms of air compressors. No skilled attendance is required, and 
the item of depreciation is merely nominal in such substantially 
erected plants as that at the Victoria Mine. By comparing the 
figures given in Tables XV and XVI, it will be seen that in the later 
installation a very marked increase was made in efficiency of 
operation ; due to improved design of the intake head, increase in 
motive head producing the flow of the compressing water, and a 
more complete separation of the air from the water in the receiv- 
ing chamber. 

It has been suggested that it might be feasible to employ the 
system in connection with an ordinary compressor plant. That is, 
to produce a low air pressure by the water plant, and then to ad- 
mit this air to the compressor cylinder where it would be brought 
up to the required higher tension. In effect, this would 
be stage compression, in which the air would be completely 

* Since the Victoria plant was put in operation, trouble has been experienced 
by the freezing up of the small pipes of the intake heads, due to the severe 
winter climate of the region. This has led to the removal of the heads, as origi- 
nally designed, the water being allowed simply to flow into the top of the com- 
pressing pipes. I am informed that the capacity of the plant, in cubic feet of free 
air compressed per minute, is practically the same as when the intake heads were 
in use (May, 1909). 

t The last-mentioned is described in Compressed Air, April, 1906, p. 3,980. 



246 COMPRESSED AIR PLANT 

cooled to normal temperature before entering the high-pressure 
cylinder. 

In 1907-8 an underground hydraulic air compressor was 
installed at one of the silver mines of Clausthal, Germany. Fig. 
121 shows, in plan and elevation, the general design of the plant, 
with details of the intake head and compressing chamber. A 
flow of water in the tunnel / is led through an 8J-inch cast-iron 
pipe, a, to the vertical air intake, b. This, shown in longitudinal 
section in the detail cut, consists of a number of flaring cast-iron 
rings, /, in the upper rim of each of which is a series of small holes, 
k, for admitting the air. Additional inlet area is provided at the 
top of the intake by a nest of small curved pipes, m. The air is 
thus drawn into the pipe and entrained by the downward flow of 
the water. The mixed air and water pass into the 8J-inch 
compression pipe, c, 492 ft. long. This pipe is laid in an inclined 
shaft, and discharges into the bottom of the compressing chamber, 
d, shown also in detail. The chamber is 52 in. diameter by 14 
ft. 9 in. high. From a point near its top the compressed air passes 
through the pipe n to the automatic check-valve e, and thence, by 
the pipe h, to the receiver i. Pipe p conveys the air from the 
receiver to the working places of the mine. The water leaves the 
compressing chamber by the 8J-inch pipe /, which discharges at 
a point 164 ft. above, into a tail-race occupying the mine level u. 
An equalizing discharge pipe, g, from the compressing chamber, 
is led up the shaft, parallel to /, entering the latter at the level of 
the tail-race. The total cost of the plant, installed, is stated to 
be $3,750.* 

The average flow of water is 792 gallons per minute; which, 
falling through a vertical height of 325 ft. (from intake to discharge 
at the tail-race), produces theoretically 66.3 horse-power. In 
testing the plant, the water was measured by a weir and the 
quantity of air compressed by filling a receiver of known capacity. 
It was found that a flow of 845 gallons per minute gave 353 cu. ft. 
of air, at a gauge pressure of 71.2 lbs. To compress i cu. ft. of 

* Abstracted from a description by P. Bernstein, in Gluckauj, March 14, 1908. 
Translation by E. K. Judd in Engineering and Mining Journal^ August i, 1908, p. 
228. 




wwjw^w»a;.Ji p «W:JWj > J<<.«'.w^ ' -w<^^, 



4 k 



AIR COMPRESSION BY ACTION OF FALLING WATER 247 

air adiabatically to this pressure requires 0.147 horse-power and 
to compress 353 cu. ft., about 51.9 horse-power. Since 70.5 
theoretical horse-power is produced by the flow of 845 gallons 

per minute, the efficiency is = 73.6 per cent. 

70-5 



Part Second 



TRANSMISSION AND USE OF 
COMPRESSED AIR 



CHAPTER XVI 

CONVEYANCE OF COMPRESSED AIR IN PIPES 

Certain losses due to friction take place in conveying com- 
pressed air through lines of piping. The diameter of the pipe is 
of vital importance, and when proportioned properly to the 
volume of air, and to the distance, these transmission losses 
are very small as compared with the other losses incident upon air 
compression. With the possible exception of electricity, no other 
means of power transmission can compare in efficiency with 
compressed air. The transmission losses appear in two ways: 
as loss of power, and as loss of pressure or head, indicated by 
difference in gauge reading at the ends of the line. Between these 
two losses there is a clear distinction. 

Loss of Power. The large and, to a great extent, unavoidable 
loss of power due to the heating of the ai? during compression 
and its subsequent cooling after leaving the compressor, has 
already been considered. But this cooling takes place so quickly 
in the receiver and piping that the resulting loss is not properly 
chargeable to transmission. The air assumes the temperature of 
the surrounding atmosphere in the first few hundred feet, so that 
when conveyed to long distances the calculation for transmission 
loss may be made without regard to the effect of temperature upon 
the volume of the air. In other words, the volume is taken simply 
as proportional to the absolute temperature, in atmospheres^ 

248 



I 



CONVEYANCE OF COMPRESSED AIR IN PIPES 249 

The power residing in the compressed air is due not only to its 
pressure, but also to its volume, in terms of number of cubic feet 
of free air (i.e., air at atmospheric pressure). Thus, while the 
pressure is reduced by frictional loss in transmission, yet this 
reduction in pressure is accompanied by a proportionate increase 
in volume, and a certain compensation is produced. Although the 
pressure of the air at the motor is diminished, there is no loss in 
the final volume of free air. As will be shown below, the loss of 
pressure due to the conveyance of air in pipes is small, but the 
actual loss of power is still smaller. The pipe itself acts in a 
measure like a receiver — as a reservoir of power. It is probable 
that much of the transmission power loss experienced in practice 
is due to leakage from joints and flaws in the pipe. 

Loss of Pressure or Head. For short distances the loss of 
pressure may be considered as taking place according to the laws 
governing the flow of all fluids, varying directly as the length of 
pipe, directly as the square of the velocity, and inversely as the 
diameter of the pipe. But for long distances the application of 
these laws becomes somewhat complex. In addition to the fac- 
tors just given, it is necessary to take into account the volume 
and pressure of the air, and the difference between the pressures 
at the receiver and at the end of the pipe line. All are more or 
less interdependent. A statement of the case, more accurate 
than the above, is as follows : For a given diameter of pipe, when 
the volume of compressed air discharged and its initial pressure 
remain constant, the loss of pressure is proportionate to the 
length of the pipe. 

But in actual service the initial pressure and the volume of 
discharge do not remain constant, and, in the passage of the air 
through the pipe, other modifying factors must be taken into 
account. In flowing through a long line of piping the pressure is 
gradually reduced by friction, while the volume is correspondingly 
increased. Therefore, to maintain in the pipe the flow of a 
given quantity of air whose volume is constantly increasing, the 
velocity also must increase, and this requires an increase of head 
or pressure. 



250 COMPRESSED AIR PLANT 

The formulas commonly used are constructed on the hy- 
pothesis that the loss of head is proportional to the length of pipe, 
so that, if a certain head be required to maintain the flow of a 
given quantity of air in a pipe 1,000 feet long, twice this head 
would suffice for a pipe 2,000 feet long. But in this case, when 
the air has passed through the first thousand feet of pipe its mo- 
tive head has been lost; and as the volume has thereby increased, 
a greater head will be necessary to maintain the flow in the second 
thousand feet. In other words, the ordinary formulas do not 
take into account the increase of volume due to the reduction 
of pressure, i.e., loss of head. 

To transmit a given volume of air at a uniform velocity and 
loss of pressure it would be necessary to construct the pipe with a 
gradually increasing area. This of course is impracticable, and 
if the rate of discharge is to be kept constant in pipe of uniform 
section, both volume and velocity must increase as the pressure is 
reduced by friction. The loss of head in properly proportioned 
pipes is so small, however, that in practice the increase in volume 
is usually neglected. 

The actual discharge capacity of piping is not proportional 
to the cross-sectional area alone — that is, to the square of the 
diameter. Although the periphery is directly proportional to the 
diameter, the interior surface resistance is much greater in a small 
than in a large pipe, because as the pipe becomes smaller the 
ratio of perimeter to area increases. To pass a given volume 
of compressed air a i-in. pipe of given length requires over 
3 times as much head as a 2-in. pipe of the same length. The 
character of the pipe also, and the condition of its inner sur- 
face, have much to do with the friction developed by the flow 
of air. Besides imperfections in the surface of the metal, the 
irregularities incident upon coupling together the lengths of pipe 
must increase friction. 

There are so few reliable data that the influences by which 
the values of some of the factors may be modified are not fully 
understood, and owing to these uncertain conditions the results 
obtained from formulas are only approximately correct. Among 



CONVEYANCE OF COMPRESSED AIR IN PIPES 



251 



the formulas in common use for determining the loss of pressure 
in pipes perhaps the most satisfactory is that of D'Arcy. As 
adapted for compressed-air transmission it takes the form : 



D 



V 



d' 



-^^ or D = ^ 



\^d^ 



X 



\PlZl 



in which 

D=the volume of compressed air in cubic feet per minute 
discharged at the final pressure, 

c = a coefficient varying with the diameter of the pipe, as de- 
termined by experiment, 

d = diameter of pipe in inches,* 

I = length of pipe in feet, 

pi = initial gauge pressure in pounds per square inch, 

p2 = final gauge pressure in pounds per square inch, 

Wi = the density of the air, or its weight in pounds per cubic 
foot, at the initial pressure p^. 

The second form of the formula, as given above, will be found 
convenient for most calculations, as the factors can be considered 
in groups. 

In the following table are given the values of c, d^, and c ^/d^. 
The values of c show some apparent discrepancy for sizes of pipe 

Table XVII 



Diameter of Pipe, 


Values of 


Fifth Powers of 


Values of 


Inches. 


' 


d 


c\/"d^ 


I 


45-3 


I 


45-3 


2 


52.6 


32 


297 


3 


56-5 


243 


876 


4 


58.0 


1024 


1856 


5 


59-0 


3125 


3298 


6 


59-8 


7776 


5273 


7 


60.3 


16807 


7817 


8 


60.7 


32768 


10988 


9 


61.0 


59049 


14812 


10 


61.2 


I 00000 


19480 


II 


61.8 


161051 


24800 


12 


62.0 


248832 


30926 



* The actual diameters of wrought-iron pipe are not the same as the nominal 
diameters for all sizes. This difference is small, however, except in the i^-in. and 
i^-in. sizes, the actual diameters of which are 1.38 ins. and 1.61 ins. respectively. 



252 



COMPRESSED AIR PLANT 



larger than nine inches, but there would be no very material 
differences in the results. 

Table XVIII gives the values of w^ for initial gauge pressures 
up to loo pounds per square inch: 

Table XVIII 



Gauge Press- 
ure, Pounds. 


^1 


■ 


Gauge Press- 
ure, Pounds. 


w. 


y'zf, 


o 


0.0761 


0.276 


55 


0.3607 


0.600 


5 


0.1020 


0.319 


60 


0.3866 


0.622 


lO 


0.1278 


0-358 


65 


0.4125 


0.642 


15 


0-1537 


0.392 


70 


0.4383 


0.662 


20 


0.1796 


0.424 


75 


0.4642 


0.681 


25 


0.2055 


0-453 


80 


0.4901 


0.700 


30 


0.2313 


0.481 


85 


0.5160 


0.718 


35 


0.2572 


0-507 


90 


0.5418 


0.736 


40 


0.2831 


0-532 


95 


0.5677 


0.753 


45 


0.3090 


0-556 


100 


0.5936 


0.770 


50 


0-3348 


0-578 









To facilitate computations in connection with D'Arcy's 
formula, Table XIX has been compiled by Mr.. William 

Cox. It gives the values of ^1— — — for terminal gauge pressures 

of from 20 to 100 lbs., and for pressure losses of from i to 10 lbs.* 
Intermediate values can be obtained by interpolation. No 

allowance is made for pipe leakage, nor for incidental friction due 

to bends in the pipe. 

By using these tables all ordinary problems involved in 

compressed-air transmission can be readily solved. For example, 

given a 5 -in. pipe, 2,500 ft. long; how many cubic feet of air per 

minute at an initial pressure of 70 lbs. can be transmitted, with 

a loss of pressure of not more than 3 lbs. ? 



From Table XVII, cVd' = 3,298; from Table XIX^ 






= 2.570 and V/ = 50. Substituting in the formula already given 



D 



3^298 
50 



2.570 = 169.5 c^- ft. compressed air per minute. 



* Reproduced by permission from Compressed Air, Feb., 1898, pp. 374-376. 



CONVEYANCE OF COMPRESSED AIR IN PIPES 

Table XIX 
Values of 



253 






Final 


Losses of Pressure, p^—p^. 


Press- 




ure 
^2. lbs. 


lib. 


2 lbs. 


3 lbs. 


4 lbs. 


5 lbs. 


6 lbs. 


7 lbs. 


8 lbs. 


gibs. 


10 lbs. 


20 


2.325 


3.241 


3.918 


4.466 


4.930 


5.336 


5.693 


6.014 


6.309 


6.574 


21 


2.293 


3.198 


3.868 


4.410 


4. 


870 


5.272 


5- 


627 


5.946 


6-237 


6. 


502 


22 


2.262 


3-157 


3.819 


4.356 


4- 


812 


5. 211 


5- 


564 


5-878 


6.168 


6. 


432 


23 


2.233 


3.117 


3.772 


4.304 


4- 


756 


5.152 


5- 


501 


5.814 


6.102 


6. 


362 


24 


2.205 


3.079 


3.727 


4-254 


4- 


702 


5-093 


5- 


440 


5.752 


6.036 


6. 


296 


25 


2.178 


3.042 


3.684 


4.206 


4- 


649 


5-036 


5- 


381 


5.688 


5-973 


6. 


233 


26 


2.152 


3.007 


3.642 


4.158 


4 


597 


4-981 


5- 


323 


5.630 


5-913 


6. 


173 


27 


2.127 


2-973 


3.601 


4. 112 


4 


548 


4.928 


5- 


268 


5.572 


5-856 


6. 


113 


28 


2.103 


2.939 


3-561 


4.068 


4 


499 


4-877 


5- 


215 


5.518 


5-799 


6. 


056 


29 


2.079 


2.907 


3-523 


4.024 


4 


452 


4.828 


5. 


164 


5.466 


5.745 


5- 


999 


30 


2.056 


2.876 


3.485 


3.982 


4 


408 


4-781 


5- 


114 


5-414 


5.691 


5- 


942 


31 


2.034 


2.844 


3.448 


3-942 


4 


365 


4-735 


5- 


066 


5-364 


5-637 


5- 


888 


32 


2.012 


2.815 


3.414 


3-904 


4 


323 


4.690 


5- 


019 


5-312 


5-586 


5- 


834 


Z2> 


1. 991 


2.786 


3-381 


3.866 


4 


282 


4.646 


4. 


971 


5-264 


5.535 


5 


782 


34 


1. 971 


2.759 


3-348 


3.830 


4 


242 


4.603 


4- 


926 


5.216 


5.487 


5 


733 


35 


1.952 


2.733 


3.317 


3-794 


4 


202 


4.561 


4 


881 


5-170 


5.439 


5 


686 


36 


1.933 


2.707 


3.286 


3.758 


4 


164 


4.520 


4 


839 


5.126 


5.394 


5 


639 


37 . 


1. 915 


2.682 


3.255 


3.724 


4 


126 


4.480 


4 


797 


5.084 


5-349 


5 


594 


38 


1.897 


2.656 


3.225 


3.690 


4 


090 


4.441 


4 


757 


5.042 


5.307 


5 


550 


39 


1.879 


2.632 


3.196 


3.658 


4 


054 


4.404 


4 


717 


5.002 


5.265 


5 


509 


40 


1.862 


2.608 


3.168 


3.626 


4 


020 


4.368 


4 


680 


4.962 


5.226 


5 


468 


41 


1.845 


2.585 


3.140 


3-596 


3 


987 


4.333 


4 


643 


4.924 


5.187 


5 


426 


42 


1.829 


2.563 


3. 114 


3.566 


3 


956 


4.299 


4 


609 


4.888 


5.148 


5 


385 


43 


1. 813 


2.542 


3.088 


3.538 


3 


924 


4.267 


4 


575 


4.852 


5.109 


5 


•344 


44 


1.798 


2.521 


3.064 


3.510 


3 


■895 


4.235 


4 


.540 


4-814 


5.070 


5 


•306 


45 


1.783 


2.501 


3.040 


3.484 


3 


.866 


4.203 


4 


.506 


4-778 


5.034 


5 


.268 


46 


1.769 


2.481 


3.017 


3.458 


3 


■837 


4. 171 


4 


.471 


4-744 


4.998 


5 


.230 


47 


1.755 


2.462 


2.995 


3.432 


3 


.808 


4-139 


4 


•439 


4.710 


4.962 


5 


.192 


48 


1.742 


2.444 


2.972 


3.406 


3 


•779 


4.109 


4 


.408 


4.676 


4.926 


5 


-155 


49 


1.729 


2.426 


2.950 


3.380 


3 


-75-' 


4-cSo 


4 


.376 


4.642 


4.890 


5 


.120 


50 


1. 716 


2.407 


2.927 


3-356 


3 


-725 


4.051 


4 


.344 


4.608 


4-857 


5 


-085 


51 


1.703 


2.389 


2.906 


3.332 


3 


.698 


4.022 


4 


■3-^3 


4-578 


4-824 


5 


.050 


52 


1.690 


2.372 


2.886 


3.308 


3 


.671 


3.993 


4 


-283 


4.546 


4.791 


'5 


.015 


53 


1.678 


2.355 


2.865 


3.284 


3 


.645 


3-965 


4 


-254 


4.516 


4-758 


4 


-983 


54 


1.666 


2.338 


2.844 


3.260 


3 


.620 


3-938 


4 


-225 


4.484 


4.728 


4 


-952 


55 


1.654 


2.321 


2.823 


3-238 


3 


-596 


3.Q11 


4 


.196 


4-456 


4.698 


4 


.920 


56 


1.642 


2.304 


2.804 


3.216 


3 


■571 


3-885 


4 


.169 


4.428 


4-668 


4 


.889 


57 


1.630 


2.289 


2.785 


3.194 


3 


•547 


3.860 


4 


-143 


4.400 


4-638 


4 


.860 


58 


1. 619 


2.273 


2.766 


3.172 


3 


-524 


3.835 


4 


.117 


4-372 


4. 611 


4 


.832 


59 


1.608 


2.258 


2.747 


3-152 


3 


502 


3. 811 


4 


.091 


4-346 


4-584 


4 


.803 


60 


1.597 


2.242 


2.730 


3.132 


3 


479 


3-787 


4 


.066 


4.320 


4-557 


4 


.775 


61 


1.586 


2.228 


2.712 


3. 112 


3 


458 


3-764 


4 


042 


4-294 


4.530 


4 


.747 


62 


1.576 


2.214 


2.695 


3.092 


3 


437 


3.742 


4 


019 


4.268 


4.503 


4 


.718 


63 


1.566 


2.200 


2.678 


3.074 


3 


417 


3.720 


3 


995 


4.244 


4.476 


4 


-693 


64 


1.556 


2.186 


2.662 


3.056 


3 


397 


3-698 


3 


971 


4.220 


4.452 


4 


.668 


65 


1.546 


2.173 


2.647 


3-038 


3 


376 


3.676 


3 


948 


4. 196', 4.428 


4 


.642 


66 


1.537 


2.160 


2.631 


3.020 


3 


356 


3.654 


3 


926 


4.172 


4.404 


4 


.617 



254 



COMPRESSED AIR PLANT 



Table XIX — Continued 



Values of 



^ 



p-p, 



Final 




Losses of Pressure, pi—p^. 


Press- 






ure, ~ 






















P^ lbs. 3 


lb. 


2 lbs. 


3 lbs. 


4 lbs. 


5 lbs. 


6 lbs. 


7 lbs. 


8 lbs. 


gibs. 


10 lbs. 


67 I 


.528 


2.147 


2.615 


3.002 


3-337 


3-634 


3-905 


4.150 


4-380 


4-592 


68 I 


•519 


2.134 


2.600 


2.984 


3-318 


3-615 


3.884 


4.128 


4-356 


4-566 


69 I 


510 


2. 122 


2.584 


2.968 


3-300 


3-596 


3-863 


4.104 


4-332 


4-541 


70 I 


501 


2.100 


2.570 


2.952 


3-283 


3-576 


3.842 


4.082 


4-308 


4-516 


71 I 


492 


2.098 


2-556 


2.936 


3-265 


3-556 


3.820 


4.060 


4.284 


4-494 


72 I 


484 


2.086 


2-543 


2.920 


3-247 


3-537 


3-799 


4.038 


4.263 


4-471 


73 I 


476 


2.075 


2.529 


2.904 


3.229 


3.517 


3-778 


4.018 


4.242 


4.449 


74 I 


468 


2.064 


2-515 


2.888 


3. 211 


3-498 


3-759 


3-998 


4.221 


4-427 


75 I 


460 


2.052 


2.501 


2.872 


3-193 


3.480 


3-741 


3-978 


4.200 


4-405 


76 I 


452 


2.041 


2-487 


2.856 


3-177 


3-463 


3-723 


3-958 


4.179 


4-383 


77 I 


444 


2.030 


2.473 


2.842 


3.162 


3-446 


3-704 


3-938 


4.158 


4.361 


78 I 


436 


2.019 


2.461 


2.828 


3.146 


3-429 


3.686 


3.918 


4-137 


4-339 


79 I 


428 


2.009 


2.449 


2.814 


3-130 


3-412 


3-667 


3.898 


4. 116 


4-317 


80 I 


421 


1.999 


2-437 


2.800 


3-115 


3-395 


3.648 


3.878 


4-095 


4-294 


81 I 


414 


1.989 


2.425 


2.786 


3.009 


3-377 


3-630 


3-858 


4.074 


4-272 


82 I 


407 


1.979 


2.413 


2.772 


3.084 


3-360 


3. 611 


3.840 


4.053 


4-253 


!3 ' 


400 


1.969 


2.401 


2.758 


3.068 


3-343 


3-593 


3.820 


4-035 


4-234 


84 I 


393 


1-959 


2.388 


2.744 


3-052 


3-326 


3-575 


3.802 


4.017 


4.215 


85 I 


386 


1.949 


2-376 


2.730 


3-037 


3-310 


3-559 


3.786 


3.999 


4.196 


86 I 


379 


1-939 


2.364 


2.716 


3.022 


3-294 


3-543 


3-768 


3.981 


4.177 


87 I 


372 


1.929 


2-352 


2.702 


3.008 


3-279 


3-527 


3-752 


3-963 


4.158 


88 I 


365 


1.920 


2.340 


2.690 


2.994 


3-265 


3-511 


3-734 


3-945 


4.139 


89 I 


358 


1. 910 


2.330 


2.678 


2.981 


3-250 


3-495 


3-718 


3-927 


4.120 


90 I 


351 


1. 901 


2.319 


2.666 


2.967 


3-235 


3-479 


3-700 


3-909 


4.101 


91 I 


345 


1.893 


2.309 


2.654 


2.954 


3-221 


3-463 


3.684 


3.891 


4.082 


92 I 


339 


1.884 


2.298 


2.642 


2.940 


3.206 


3-447 


3.666 


3-873 


4.064 


93 I 


333 


1.876 


2.288 


2.630 


2.927 


3-191 


3-432 


3-650 


3-855 


4.048 


94 I 


327 


1.867 


2.278 


2.618 


2.914 


3-177 


3.416 


3-634 


3-840 


4.032 


95 I 


321 


1.859 


2.267 


2.606 


2.900 


3.162 


3.401 


3.618 


3-825 


4.016 


96 I 


315 


1.850 


2.257 


2-594 


2.887 


3.148 


3-387 


3-604 


3.810 


4.000 


97 I 


309 


1.842 


2.246 


2.582 


2-873 


3-135 


3-373 


3-590 


3-795 


3-984 


98 I 


303 


1-833 


2.236 


2.570 


2.862 


3-123 


3-360 


3-576 


3.780 


3-969 


99 I 


297 


1.825 


2.226 


2.560 


2.851 


3. no 


3-347 


3-562 


3-765 


3-953 


100 I 


291 


1. 817 


2.217 


2-550 


2.840 


3.098 


3-334 


3-548 


3-750 


3-937 



Volumes of compressed air are easily converted into corre- 
sponding volumes of free air by multiplying by the absolute press- 
ure in terms of atmospheres (i atmosphere = 14. 7 lbs.). Thus, 
100 cu. ft. of air at 80 lbs. gauge pressure, or 94.7 absolute 
pressure, are equal to 644 cu. ft. of free air, at sea-level. 
Table XIII gives the air pressures in pounds per square inch for 



CONVEYANCE OF COMPRESSED AIR IN PIPES 255 

altitudes up to 15,000 ft., with the corresponding barometric 
readings. 

Another formula for the loss of pressure in pipes has been 
published by Mr. Frank Richards, as follows : * 

10,000 Da 

D= diameter of pipe in inches. 

L = length of pipe in feet. 

V = volume of compressed air delivered, in cubic feet per 
minute. 

H = head or difference of pressure required to overcome fric- 
tion and maintain the flow. 

a = constant for diameter of pipe. 

Values of a for Different Nominal Diameters 
Of Wrought -Iron Pipe. 



I'' . 


. . 0-350 


f . 


• . 0.730 


5". 


. . 0.934 


ir. 


. . o.soof 


3r. . 


. 0.787 


6". 


. I . 000 


ir. 


. . o.662t 


4" . 


. 0.840 


8''. 


. . 1-125 


2" . 


. . 0.565 






10''. 


. . 1.200 


2r. 


. . 0.650 






12" . . 


. 1.260 



Using this formula with its constants, the calculated losses of 
pressure are smaller, and, conversely, the volumes of air discharged 
are larger, under the same conditions, than those obtained from 
D'Arcy's formula. 

The losses of pressure in a table by F. A. Halsey indicate that 
the constants used by him differ materially from those given 
above. For comparison a series of random examples are shown 
in Table XX. 

An examination of this table shows that in all cases the figures 
from D'Arcy's formula he between the others, and until further 
experimental data are available it would appear safe to conclude 
that the results obtained from this formula are sufficiently ac- 

* American Machinist, Dec. 27th, 1894, 

t The values of a for i^- and ij-in. pipe are not consistent with those for 
other sizes. See foot-note on page 251. 



256 



COMPRESSED AIR PLANT 

Table XX 



Cubic Feet of 


Length of Pipe, 
Feet. 


Diameter of 
Pipe, Inches. 


Transmission losses, Pounds. 


Pressure. 














Richards. 


D'Arcy(Cox). 


Halsey. 


1,000 


1,000 


4 


3-^3 


3-71 


5.02 


1,000 


1,000 


5 


■95 


1. 17 


1.63 


1,000 


1,000 


6 


-35 


-46 


.64 


4,000 


5,000 


8 


5-92 


8-44 


13-05 


4,000 


5,000 


10 


1.78 


2.81 


4.20 


4,000 


5,000 


12 


.68 


1.06 


1.70 



curate for ordinary calculations. It must be remembered that, 
within certain limits, the loss of head or pressure increases with 
the square of the velocity. To obtain the best results it is found 
in practice that the velocity of flow in the main air pipes should 
not exceed twenty or twenty-five feet per second. Experiments 
made to determine the loss of pressure in the mains of the Paris 
compressed-air plant gave the following results : * 

Table XXI 
Djameter of Pipe, Twelve Inches 



Velocity of Flow in 
Feet per Second. 


Initial Pressure, 
Pounds. 


Final Pressure, 
Pounds. 


Per Cent, of Initial 
Pressure Lost per Mile. 


25 

50 

100 


100 
100 
100 


97-6 
90.6 
53-8 


2-4 

9-4 

46.2 



It is evident that when the initial velocity much exceeds 50 
ft. per second the percentage loss becomes very large; and, fur- 
thermore, by using piping large enough to keep down the velocity 
the friction loss may be almost eliminated. For example, at the 
Hoosac tunnel, in transmitting 875 cu. ft. of free air per 
minute at an initial pressure of 60 lbs., through an 8-in. pipe 
7,150 ft. long, the average loss including leakage was only 2 lbs. 
The velocity in this case was 8J ft. per second. A volume of 
500 cu. ft. of free air per minute, at 75 lbs. gauge pressure, can 

* Unwin. Van Nostrand's Science Series, No. 106, p. 78. 



I 



CONVEYANCE OF COMPRESSED AIR IN PIPES 257 

be transmitted through 1,000 ft. of 3-in. pipe with a loss of 4.1 
lbs., while if a 5-in. pipe were used the loss would be reduced to 
.24 lb., the velocities being respectively 28 ft. and 10 feet per sec- 
ond. In driving the Jeddo mining tunnel, at Ebervale, Luzerne 
Co., Penna., two 3j-in. machine drills were used in each heading, 
with a 6-in. main, the maximum distance of transmission being 
about 10,800 ft. This pipe was so large in proportion to the vol- 
ume of air required for the 2 drills (about 230 cu. ft. free air per 
minute) that the loss was reduced to an extremely small quan- 
tity, the velocity being only 3 J ft. per second. A calculation 
shows a loss of .002 lb., and the gauges at each end of the 
main were found to record practically the same pressure. 

A due regard for economy in installation, however, must limit 
the use of very large piping, the cost of which should be considered 
in relation to the cost of air compression in any given case. Diam- 
eters of from 4 to 6 ins. for the air mains are large enough for 
operating simultaneously from 6 to 10 drills. Up to a length 
of 3,000 ft. a 4-in. pipe will carry per minute 480 cu. ft. of free 
air compressed to 82 lbs., with a loss of 2 lbs. pressure. This vol- 
ume of air will run four 3-in. drills. Under the same conditions 
a 6-in. pipe, 5,000 ft. long, will carry 1,100 cu. ft. of free air per 
minute, or enough for 10 drills in constant operation. 

A mistake is often made in putting in branch pipes of too 
small a diameter. For a distance of, say, 100 ft. a i J-in. pipe is 
small enough for a single drill, though i-in. is frequently used. 
While it is, of course, admissible to increase the velocity of 
flow in short branches considerably beyond 20 ft. per second, 
extremes should be avoided. To run a 3-in. drill from a i-in. 
pipe 100 ft. long would require a velocity of flow of about 55 ft. 
per second, causing a loss of 10 lbs. pressure. In this connection 
Table XXII * may be studied with advantage. 

Compressed-Air Piping. The pipe for conveying compressed 
air may be of cast or wrought iron. If of wrought iron, as is 
customary, the lengths are connected either by sleeve couplings 
or by cast-iron flanges into which the ends of the pipe are ex- 

* From the catalogue of the Norwalk Iron Works Co. 



258 



COMPRESSED AIR PLANT 

Table XXII 



Nomina 
of Pipe, t 


I Size 


I 


in. 


ij in. 


I2 in. 


2 ins. 


2j ins. 


1 
1 


Length of Pipe 
in Feet. !il^=- 


50 


100 


100 


300 


100 


300 


200 


500 


250 


600 


r/) 




W 
> 

Q 

O 
H 

O 
Ah 

M 

H 

H 

W 
P^ 


79.8 


23.2 


16.4 


35-2 


20.3 


63-6 


36-7 


84-7 


53-6 


142. 


91.7 


P^ 
(fi 




79-6 
79-4 


33--^ 


23-4 


49-7 


28.7 


89-9 


51-9 


119.^ 


75-7 


200.9 


129.6 




40.4 


28.6 


61.0 


35-2 


109. 1 


63.0 


146.5 


92-7 


244.4 


157-7 


P^ 


W 


79-2 


46.8 


2>?>-^ 


70-3 


40.6 


127. 1 


73-4 


169. 1 


107. 1 


283.2 


183. 1 






79- 


52-3 


37-° 


78.6 


45-4 


142.0 


82.0 


189. 1 


119. 7 


317-1 


204.6 


< 
H 




78.8 


57-1 


40.4 


86.1 


49-7 


155-4 


89-7 


207. 


131. 


348.4 


224.8 




PQ 
W 


78.6 
78.4 


61.6 


43-6 


93-0 


53-7 


168.0 


97.0 


223.3 


141-3 


377-0 


243-9 


H 

H 

< 

P^ 


65-9 


46.6 


99-2 


57-3 


179-3 


103-5 


238.7 


151. 1 399.6 


258.4 


Ph 


78.2 


70-3 


49-7 


105.4 


60.8 


190.5 


IIO.O 


252.9 


160. 1 424.1 


273.6 


< 
Q 




78. 


73-7 


52.1 


no. 8 


64.0 


200.7 


115-9 


266.5 


168.7 


446.7 


288.6 


tn 
m 
W 
Pi 
P4 




o 


77-8 


77-2 


54-6 


116. 2 


67.1 


209.9 


121. 2 


279.2 


176.7 


469.0 


302.6 




77-6 


80.7 


57-1 


121. 4 


70.1 


219. 1 


126.5 


291.5 


184.5 


489.6 


315.9 


< 

H 

w 


77-4 


84.0 


59-4 


126.3 


72.9 


228.1 


131-7 


303.4 


192.0 


509.3 


328.6 


in 

< 


77-2 


87.1 


61.6 


131. 1 


75-7 


236.7 


136.6 


314.4 


199.0 


528.3 


340.8 


04 


w 

H 

H 


77- 


90-3 


63-7 


135-4 


78.2 


245.2 


141. 6 


325.5 


206-0 


546.5 


352.6 


> 


76.8 


92-9 


65-7 


139.8 


80.7 


252.4 


145-7 


336-1 


212.7 


564.2 


364.0 


a 

I— 1 
< 

P4 

^■ 


H 

a 

s 

a 




76.6 


95-6 


67-7 


143-9 


83-1 


259-8 


150.0 


346.2 


219. 1 


581.3 


375.0 




76-4 


98-4 


69.6 


148. 1 


85-5 


267.6 


154.5 


356.0 


225.3 


597-5 


385.5 


w 


76.2 


lOI.O 


71-5 


152.1 


87.8 


274.7 


158.7 


365.6 


231.4 


613.8 


396.0 




76. 


103.8 


73-4 


156. 1 


90.1 


281.3 


162.4 


375.6 


237.3 


629.3 


406.0 


t-H 


75-8 


106.3 


75-2 


159-7 


92.2 


288.4 


166.6 


383.9 


243.0 


644-5 


415-8 


75.6 


108.7 


76.9 


163-3 


94-3 


295-5 


170.6 


392.8 


248.6 


659.2 


+25.3 




75-4I 


III.O 


78.5 


167.0 


96.4 


301.7 


174-2 


401.4 


254-0 


673-8 


434-7 




75-2 


II3-3 


80.1 


170.4 


98.4 


307-9 


177.8 


409.7 


259-3 


687.8 


443-8 




75- 


II5-5 


81.7 


173-9 


100.4 


314.3 


181. 5 


417.9 


264.5 


701.6 


452.7 



I 



CONVEYANCE OF COMPRESSED AIR IN PIPES 259 

panded or screwed. Sleeve couplings are used for all except the 
large sizes. The smaller sizes, up to ij in. are butt-welded, 
while all from i J in. up are lap-welded to insure the necessary 
strength. Extra heavy piping may be had for higher pressures 
than those commonly used. Wrought-iron spiral-seam riveted, 
or spiral-weld steel, tubing is sometimes used. It is made in 
lengths of 20 ft. or less. For convenience of transport in re- 
mote regions rolled sheets in short lengths may be had. They 
are punched around the edges, ready for riveting, and are packed 
closely, 4, 6 or more sheets in a bundle. 

All joints in air mains and branches should be carefully made. 
The pipe may be tested from time to time by allowing the air at 
full pressure to remain in the pipe long enough to observe the 
gauge. In case a leak is indicated it should be traced and stopped 
immediately. Air leaks are more expensive than steam leaks 
because of the losses already suffered in compressing the air. In 
putting together screw joints care should be taken that none 0} 
the white lead or other cementing material is forced into the pipe 
This would cause obstruction and increase the friction loss. Also,, 
each length as put in place should be cleaned thoroughly of all 
foreign substances which may have lodged inside. To render 
the piping readily accessible for inspection and stoppage of leaks 
it should, if buried, be carried in boxes sunk just below the sur- 
face of the ground; or, if underground, it should be supported 
upon brackets along the side of the mine workings. Low points 
in pipe lines, which would form " pockets " for the accumulation of 
entrained water, should be avoided, as they obstruct the passage 
of the air. In long pipe lines, where a uniform grade is im- 
practicable, provision may be made near the end for blowing out 
the water at intervals, when the air is to be used for pumps, 
hoists, or other stationary engines. 

For long lengths of piping expansion joints are required, par- 
ticularly when on the surface. Underground they are not often 
necessary, as the temperature is usually nearly constant, except in 
shafts, or elsewhere, where there may be considerable variations 
of temperature between summer and winter. 



26o 



COMPRESSED AIR PLANT 



As each bend or elbow in a pipe line has a serious retarding 
effect, abrupt changes in direction and sharp curves should be 
avoided so far as possible. For the same diameter of pipe the 
resistance caused by a bend increases as the radius of the curve 
diminishes, but the exact relation is not accurately known. In 
the absence of sufficient experimental data the following table is 
given, as published in the catalogue of the Norwalk Iron Works 
Co.: 

Table XXIII 



Radius of elbow in terms of 
diameter of pipe 


5 


3 


2 


i-i 


li 


I 


f 


y 




Equivalent length of straight 
pipe in terms of its diameter. . 


7-85 


8.24 


9-03 


10.36 


12.72 


17-51 


35-09 


121. 2 



It would appear that these allowances are none too large, 
since for steam piping the frictional resistance of each ordinary 
sharp right-angled elbow is considered equivalent to that due to 
a length of straight pipe equal to forty times its diameter. How- 
ever, in putting in wrought-iron air piping of the sizes customarily 
used the bends are not necessarily so sharp as a standard right- 
angled elbow. When many sharp bends are permitted, it is 
evident that the resistance may become very great. 

Under most conditions this difficulty may be avoided by the 
exercise of proper care in the installation of the pipe lines. The 
matter should have special consideration in the stopes of mines 
timbered with square sets. As far as possible, the piping should 
be carried diagonally through the sets, bending the pipe itself 
whenever necessary, instead of using right-angled elbows. 



CHAPTER XVII 

COMPRESSED AIR ENGINES 

Compressed air may be employed as a motive power in an 
engine in two ways, viz : at full pressure or expansively. By 
working at full pressure it is understood that the air is admitted to 
the cylinder throughout practically the entire length of stroke, i.e.^ 
without cut-off, and that therefore nearly a cylinderful of air at 
gauge pressure is exhausted at each stroke. In this case the work 
of the air engine is roughly similar to that done in anon-expansive- 
working steam engine. Among the machines which use air in 
this way are rock-drills and simple, direct-acting pumps, without 
rotary parts. 

By the term expansive-working it is meant that the air is 
admitted to the cylinder during only a part of the stroke, and is 
then cut off and the stroke completed by the expansive force of the 
air. For operating in this way some equalizing agent, such as the 
fly-wheel, is essential, and as a rule a higher initial pressure is 
employed than when w^orking under full pressure throughout 
the stroke. It is necessary to distinguish between complete and 
partial or incomplete expansion. When the air is used with com- 
plete expansion the operation in the cylinder is the reverse of 
adiabatic compression in a compressor, the final pressure being 
equal to that of the atmosphere. But as no condensation is 
possible with air, it follows that the lowest terminal pressure in 
the cylinder must still be sufficiently above atmospheric pressure 
to produce a proper exhaust, and to overcome the friction of the 
engine at the end of the stroke. Hence, theoretically complete 
expansion is impracticable for simple air engines of ordinary 
design. 

Most air engines work with partial or incomplete expansion, 

261 



262 COMPRESSED AIR PLANT 

the air expanding adiabatically in the latter part of the stroke. 
The point of cut-off is such that the terminal cylinder pressure 
exceeds the back-pressure by an amount sufficient to cause a free 
exhaust. In the conditions here set forth, no reference is made 
to the thermal changes incident upon adiabatic expansion in the 
air cylinder. Although in principle compressed air is used like 
steam, both being elastic fluids, there is an essential difference in 
the results obtained, due to the reduction in temperature. In ex- 
panding behind the piston, a given volume of compressed air at 
a given pressure will not produce the same amount of power as 
steam under the same conditions. If two curves be constructed, 
representing the expansion of equal volumes of air and steam, from 
the same initial pressure down to pressures below that of the 
atmosphere, it will be seen that the steam pressure at all points 
of the stroke is considerably higher than the air pressure; and 
the expansion curve of the air reaches the atmospheric line much 
sooner than the steam curve. 

Fig. 122 shows an ideal card, in which the initial pressure is 
75 lbs., and the cut-off is at ^ stroke. The adiabatic expansion 
curve of the air shows that the pressure is reduced to zero gauge 
pressure when the air has expanded to 3I times the initial volume, 
the mean effective pressure being 18.9 lbs. At the end of the 
stroke the pressure falls to 7 lbs. below atmospheric pressure. 
The steam curve, on the other hand, does not cut the atmospheric 
line until the expansion reaches 4J times the initial volume, and 
the mean effective pressure is 25.2 lbs. The lower mean pressure 
of the air is due to the development of cold during its expansion. 
The operation is the reverse of compression, and the resulting loss 
of motive power is analogous to the loss of work in the compressor 
caused by the generation of heat. Just as the heat of com- 
pression reacts upon the air while being compressed in the cylin- 
der, and produces a higher tension than that due to the mere 
reduction in volume; so conversely, when expansion takes place, 
the air, which is usually at normal atmospheric temperature on 
entering the cylinder, rapidly gives up its sensible heat, and the 
cold reacting upon the expanding air reduces its pressure faster 



COMPRESSED AIR ENGINES 



263 



LENGTH OF STROKE 

2 3 4 



p 


- 




— 


+ 




- 


- 






- 


- 


- 


— 


- 


- 


- 


- 


- 


- 


- 


— 


— 


— 


— 


- 


— 


- 


~/-> 








1 


























































I 


























































1 


























































H 


























































u 
















































A 










U 


























































u 


























































U 


























































\\ 


























































1 


[ 














































« 










I 


\ 






















































































































































































































































































A 














{ 


























































\ 


























































\ 




























































V 


























































\ 

























































\ 


^ 
























































\ 


























































> 


^ 


























































\ 




\ 






















































\ 




\ 






































A 


















s 




\ 






















































\ 






V 




















































> 


V 




\ 






















































\ 






\ 






















































V 






^'^-. 



















































1 


V 
























































\ 






[ 


S 


s. 


















































\ 










\ 


















































\ 










N 


^ 
















































\ 












^ 


^ 















































^ 


s. 














->. 














































"V 


^ 














^ 




^ 










































"^ 


^ 
































































--> 




- 




^ 










































































































































































- 


— 












— 


— 




— 


_v 


^ 


ilU 


n Lii 


1©: 


k 


— 


— 


— 


— 






— 




— 




— 



70 



60 



50 



40J 



20 



10 



Fig. 122. — Expansion Curves of Steam and Air. 



264 



COMPRESSED AIR PLANT 



than that which is due to the increase in volume alone. More- 
over, this behavior of compressed air is independent of the initial 
temperature, since the resulting expansion curve would be unal- 
tered. In the case of steam the initial temperature is high, and 
is reduced but little during expansion from ordinary working 
pressures down to atmospheric pressure. 

A similar comparison may be made for other initial pressures 
and ratios of cut-off. In every case the mean effective pressure 
is higher for steam than for air. It follows that, to develop the 
same amount of power in a given cylinder and with the same 
initial pressure, the cut-off must be later in the stroke with air 
than with steam. 

So low are the temperatures produced by the expansion of air,, 
from ordinary working pressures of sixty or seventy pounds down 
to atmospheric pressure, that for a long time the expansive use of 
compressed air was considered impracticable. In Table XXIV 
are given the theoretical final temperatures of the exhaust air, in 
working with complete expansion, and also at full pressure 
throughout the stroke, for different ratios of initial to final press- 
ure, together with the theoretical efficiencies. The initial tem- 
perature is taken as 68° F.* 

Table XXIV 



Ratio of 


Working With Complete Ex- 
pansion, 


Working at Full Pressure. 


Initial to 








Final Press- 










ure. 


Final Tempera- 


Theoretical Effi- 


Final Tempera- 


Theoretical Effi- 




ture. Degrees Fah. 


ciency. 


ture. Degrees Fah. 


ciency. 


2 


— 28.2 


-855 


-8.4 


.82 


3 


-76- 




806 


—34-5 


.72 


4 


—106.6 




782 


—45-7 


.67 


5 


—128.2 




768 


—54-4 


-63 


6 


—144.4 




•758 


-59-8 


.60 


7 


—158.8 




.751 


-63-4 


•57 


8 


—170.8 




.746 


—66.1 


.55 


9 


—180.6 




.742 


-68. 


.53 


10 


—189.2 


.739 


—69.7 


-51 



* M. Mallard, " Etude Theoretique sur les Machines a Air Comprime,'* p. 27. 
Robert Zahner, " Transmission of Power by Compressed Air," p. 100. 



COMPRESSED AIR ENGINES 265 

In the table it is shown that by working at full pressure 
extremely low temperatures of exhaust are avoided; but the 
efficiency of this method of using compressed air is necessarily 
much below that obtained from expansive working. It is under- 
stood that the temperatures here given are theoretical and are 
never actually reached in practice. The cold produced is modi- 
fied by several causes : (i) Some heat is transmitted from the ex- 
ternal atmosphere through the cylinder walls; (2) the re-com- 
pression of the clearance air at each stroke produces heat in the 
cylinder, to a degree that increases with the initial pressure and 
the clearance volume; and (3) the presence of even a small quan- 
tity of moisture in the air tends in some degree to raise the 
cylinder temperature. 

A few brief notes will here be given concerning the elements 
of the operation of compressed-air engines, that may be con- 
sidered more or less applicable for ordinary service, viz : working 
at full pressure, with partial expansion, or with complete expan- 
sion. Isothermal expansion may be neglected, since it involves 
the application of a sufficient degree of external heat to the air 
while doing its work in the cylinder to produce a terminal tem- 
perature equal to the initial temperature. 

I. Working at Full Pressure. This mode of using com- 
pressed air is common for engines like pumps, operating under a 
constant resistance and not provided with fly-wheels: 
Let P' =the absolute initial pressure of the air. 

V = the initial volume of air, at the pressure P', or K times 
the volume of one pound of air used per unit of time. 

T'=the absolute initial temperature of the compressed air. 

T =the absolute final temperature of the air at exhaust, on 
expanding to atmospheric pressure. 

P = pressure of the air at exhaust. 

W =foot-pounds of work done. 
From the theory of compressed air: 

R = J (C^— CJ=778 (0.2375 -0.1689) =53.37, where J is 
Joule's heat unit, and C^ and C^ are the specific heats of air at 
constant pressure and constant volume. 



266 COMPRESSED AIR PLANT 

As no work is done by the expansive force of the air originally 
produced by compression, W equals the volume of air used, V, 
multiplied by the difference between P' and P, or: W = V'(P' — P). 

KRT' 

Substituting for V^ its value, — :^7 — , as obtained from : P' V 

= KRT^ 

TvRT' P \ 

W= -^r- (P' - P) = KRT' (I - -) 

P \ 
Giving R its value, 53.37: W = 53.37 KT' (i - — j 

2. Working with Partial Expansion. The advantages of 
using compressed air in this way may be obtained from engines 
possessing fly-wheels, provided that the cut-off be not too early in 
the stroke to avoid excessive reduction of cylinder temperature, 
or else that the air be reheated before entering the cylinder. 

In this case the values of P', V, and T' are as above. From 
the point of cut-off the air expands adiabatically down to a ter-, 
minal pressure of P'' and volume V, the final temperature in the 
cylinder falling to T'\ On exhausting, the pressure, volume, and 
temperature become P, V, and T. The work done is composed 
of three parts, viz : 

W = work between the point of admission and the point of 

cut-off = F v. 
W' = work performed by expansion of the volume V from the 

point of cut-off to the end of the stroke = 778 KC^, 
(T'-T'O. 
W''= negative work due to back-pressure = — P V. 

Taking the algebraic sum of these three quantities : 
W = F V'-f778K C^(T'-T'0-PV'' 
But, as under (i): V =-^|^ and V" = ^|J^ 

Substituting these values of V and V, and for R and C, 
their numerical values of 53.37 and 0.1689: 

K [53.37 T' + 131.4 (T' -T") - 53.37 T' (|;)] 
= 53-37 K [t' + 2.46 (T' - T") - T'p] 



I 



w 



COMPRESSED AIR ENGINES 



26"] 



3. Working with Complete Expansion. In the theoretical 
card, Fig. 123, are shown the relations of the compression and ex- 
pansion lines, the shaded portion representing the useful work 
done by the complete expansion of cold air in a motor cylinder. 




When the expansion is adiabatic, the same relations exist between 
pressures, volumes, and temperatures as were set forth in the dis- 
cussion of adiabatic compression, viz : 

The theoretical work done by complete adiabatic expansion 
may be expressed by a formula similar to that employed for com- 
pression, but with an inversion of certain of the quantities, thus: 

W = ;^PV[.-(J)'^'], in which 
W = the theoretical foot-pounds of work done by the expan- 



268 COMPRESSED AIR PLANT 

sion to atmospheric pressure of i pound (13. i cu. ft.) of free air. 
Substituting the values of the constants, and for working at sea- 
level: 

W = 3.463 X 144 X 14.7 X 13.1X 

[.-(^)-]=0,o.,[.-(^')"] 

For example, if P' be 40 lbs. gauge pressure: 

W = 96,029 I I — f — —] I =30,440 ft. lbs., or 2,323 ft. lbs. 

per cu. ft. of free air. 

Actual Work Done. In the above expressions no account is 
taken of the friction of moving parts of the motor engine, or loss 
of work caused by leakage. In determining the actual work, the 
general case will be where a cut-off is employed. The relations 
between initial and terminal pressures and temperatures, for 
different ratios of expansion in a motor-engine cylinder, are 
shown in Table XXV,* the points of cut-off, in tenths of the 
cylinder stroke, being given in the first column. 

The quantities in Table XXV must be further corrected for 
piston clearance and the lost volume represented by the air ports 
and passages of the cylinder, because part of the air expands into 
these clearance sf)ace&. Therefore, the actual effect of the cut- 
off, in any given case, is found by dividing the sum of the cut-off 
plus clearance, by the cylinder volume plus clearance. For ex- 
ample, if the stroke be 10, with a cut-off of -^-q, and clearance of 6 
per cent., the actual volume of the cylinder, including clearance, 
will be: (10X.06) 4-io = io.6. Then the ratio of actual cut-off, 
plus clearance, is 4 -f. 6 =4.6, and the working cut-off becomes 
4.6 -T- 10.6 = 0.434. In this way Table XXVI has been constructed, 
for use in connection with Table XXV. It shows the actual cut- 
off corresponding to the different nominal points of cut-off, for 
the percentages of piston clearance named at the top of the col- 
umns. 

* This table, as well as Table XXVT, is taken in part from those used by Gard- 
ner D. Hiscox, in " Compressed Air, its Production, Uses and Application," 190 1, 
p. 202. 



COMPRESSED AIR ENGINES 



269 



Table XXV 
Theoretical Ratios of Pressures and Temperatures Due 
TO THE Expansion of Compressed Air in a Motor 
Cylinder. 





X ^ 


c3 S ^'^ 


i^2f^ . 


Sh 


S aiS . 




d 

6 


■lit 

^ 11 












O.IO 


10.00 


0.249 


0.166 


0.391 


0.513 


0.039 


•15 


6.67 


.348 


•233 


.460 


.578 




069 


.20 


5.00 


•436 


•295 


.518 


.627 




104 


•25 


4.00 


•515 


•353 


.568 


.669 




142 


•30 


3-33 


•585 


.408 


.612 


•705 




184 


•35 


2.86 


.647 


.460 


.652 


•737 




228 


.40 


2.50 


.706 


.510 


.688 


.767 




275 


•45 


2.22 


•757 


•558 


.722 


•794 




325 


•50 


2.00 


.802 


.604 


•754 


.818 




378 


•55 


1.81 


.842 


.649 


.784 


.841 




433 


.60 


1.67 


•877 


.692 


.812 


.862 




487 


•65 


i-'54 


.907 


•734 


•839 


.882 




545 


.70 


1-43 


•932 


•774 


.865 


.902 




605 


•75 


^■33 


•954 


.814 


.889 


.920 


.667 



Table XXVI 

Excess of Cut-off Due to Percentage of Clearance for 

THE Nominal Cut-offs in Column i. 



li 






Percentage 


OF Clearance. 






15 


•03 


.04 


.05 


.06 


.07 


.08 


.10 


.10 


0.126 


0-135 


0.143 


0-151 


0.159 


0.167 


0.182 




15 


•175 


.184 


.191 




198 


.206 


.213 


.227 




20 


.223 


.231 


.238 




245 


•252 


•259 


•273 




25 


.272 


.279 


.286 




293 


.299 


•305 


.318 




30 


•320 


•327 


•333 




340 


•346 


•352 


•364 




35 


.368 


•376 


•380 




387 


•392 


•398 


.409 




40 


.417 


•423 


.429 




434 


•439 


.444 


.455 




45 


•465 


•471 


•477 




481 


.486 


.490 


•500 




50 


•514 


•519 


•524 




528 


•533 


•537 


•546 




55 


•564 


.568 


•571 




S76 


•580 


•585 


•591 




60 


.612 


.615 


.619 




.623 


.626 


.630 


•637 




^5 


.660 


.664 


.667 




.670 


•673 


.676 


.682 




70 


.709 


.711 


.714 




.717 


.720 


.722 


•727 


•75 


•758 


.760 


.762 


.764 


.766 


.768 


•772 



270 COMPRESSED AIR PLANT 

The theoretical terminal cylinder pressure resulting from 
adiabatic expansion may be expressed by : 

P' I 

- — 7 — P, in which C = ratio of expansion = — : z rr- 

C^4°6 ' ^ pomt of cut-off 

(see column 2, Table XXV). 

For example, for a cut-off at — stroke and 65 lbs. gauge press- 
ure, the terminal pressure (above atmospheric pressure) will be: 
^5 + 't^ - 14.7 = 7.^ lbs. 

2 t- 1.406 ^ ' ' 

The volume corresponding to the nominal cut-off is increased 
by the clearance, and adds to the mean pressure. Thus, in the 
above example, assuming the clearance to be 6 per cent., the actual 
cut-off (Table XXVI) is increased from 0.4 to 0.434, of which the 

ratio, C, is = 2.3. From Table XXV, column 7, the ratio of 

•434 

initial to terminal pressure, corresponding to the actual cut-off 

of 0.434, is (by interpolation) .31 ; whence: (79.7 X 0.31) — 14.7 = 

10 lbs. terminal pressure. 

Cylinder Volume Required for a Given Power. The work per 
stroke is found by dividing the foot-pounds of work to be done 
per minute by twice the number of revolutions of the engine 
(which would be determined for any given size of engine by 
the ordinary empiric rules of practice). This is substituted, with 
the initial and final pressures, in the formula for working with 
full pressure, partial or complete expansion, as the case may be, 
which is then solved for the initial volume, V, of compressed 
air used per stroke. To the theoretical cylinder volume thus 
found, the allowance for piston clearance is added, according to 
the type of engine. The proper proportion between stroke and 
diameter of cylinder is finally determined. 

The volume of free air per minute, required for an air engine, 
per indicated horse-power and for different ratios of cut-off, are 
shown in Table XXVII, by F. C. Weber.* The figures given in 

* Compressed Air, Oct., 1896, p. 117. 



COMPRESSED AIR ENGINES 



271 



this table do not include the volume corresponding to piston 
clearance which may be found as already shown. 

Table XXVII 
Cubic Feet of Free Air per Minute Used in Motor Engine, 

Per I. H.-P. 



Point of 








Gauge Pressures, Pounds. 




Cut-off. 


































30 


40 


so 


60 


70 


80 


90 


100 


no 


125 


I 


23.3 


21.3 


20.2 


19.4 


18.8 


18.42 


18.10 


17.8 


17.62 


17.40 


i 


18.7 


17. 1 


16. 1 


15-47 


15.0 


14.6 


14-35 


14-15 


13.98 


13-78 


i 


i7-«5 


16.2 


15.2 


14-S 


14.2 


13-75 


13-47 


13.28 


13.08 


12.90 


h 


16.4 


14.5 


13-S 


12.8 


12.3 


11-93 


II. 7 


11.48 


11.30 


II. 10 


i 


17-5 


15.2 


12.9 


11.85 


11.26 


10.8 


10.5 


10.21 


10.02 


g.jS 


i 


20.6 


15.6 


13-4 


^3-3 


11-40 


10.72 


10.31 


10. 


9.75 


9.42 



In this table the air is supposed to be used without reheating, 
and at an initial temperature of 60° F. Reheating will reduce the 

T 

volume of air proportionally to the ratio 7^, where T2 = 459° + 60° 

I3 

= 519° F., or absolute temperature; and T3 = 459° P^^^ the tem- 
perature of the reheated air on entering the motor cylinder. 
Thus, if the air be reheated to 200° F., the above ratio becomes 

^-^ = 0.787, by which decimal the volume of air as found in 

659° ^ 

the table must be multiplied. 

So far as mine service is concerned, it has been customary 
to consider compressed air almost exclusively as an agent for the 
operation of rock-drills, and in view of its preponderating applica- 
tion to this use its adaptability under proper conditions to the 
driving of other machines and engines is sometimes overlooked. 
Of late years, however, with improved methods of compression 
and reheating, attention has been given to employing compressed 
air for a greater variety of service; not only underground, but 
for certain portions of the surface plant of mines as well. 
Aside from cases where the disposal of exhaust steam would be 



272 COMPRESSED AIR PLANT 

troublesome, the question is largely one of comparative loss in 
transmission and the power cost of the air. 

Although not strictly in place in this chapter, reference may 
be made to what has been called the " two-pipe system" or 
"high-range compressed-air transmission," introduced some 
years ago by Charles Cummings.* 

The machine or engine using the air makes in effect a closed 
circuit with the compressor. After the air has done its work in 
the motor cylinder, it is returned to the compressor at the pressure 
of the exhaust, through a second line of piping. The return pipe 
connects with a closed chamber at the compressor, in which the 
inlet valves are placed, thus enabling the compressor to begin its 
stroke with the cylinder filled under a considerable initial pressure. 
Then, after raising the pressure to the original point, the com- 
pressor delivers the air into the main, to be used again by the air 
engine. The actual working pressure of the air engine is. there- 
fore, the difference between the pressures in the delivery and 
return pipes. Barring leakage, the same air is thus used over and 
over, the intention being that the compressor shall put back into 
the air kept in circulation the power expended in the motor 
engine cylinder. 

Though the compressor itself is not materially different from 
the ordinary forms, the two-pipe system requires a rather com- 
plicated arrangement of piping and valves for charging the 
apparatus with air at the working pressure adopted, and for 
governing the speed and output according to the rate of con- 
sumption of air. f The advantages of the system are : a higher 
efficiency than is obtained from moderate-size compressors of the 
usual types, and less trouble from freezing at the motor engine 
by reason of the relative dryness of the air due to its higher 
tension. The efficiency increases with the pressure employed. 
In using compressed air without reheating the two-pipe system 

* Patent No. 456,941 was issued to Mr. Cummings in 1891. 

t A detailed illustrated description is given by Frank Richards in American 
Machinist, April 28th, 1898, p. 23. See also Compressed Air Magazine, Oct., 
1907, p. 4599. 



COMPRESSED AIR ENGINES 273 

is superior in principle to the ordinary mode of operating com- 
pressed-air plant. But because of the greater first cost its ad- 
vantages disappear when reheating can be adopted, and the 
single-pipe system is then found to be preferable. 

The two-pipe system is best suited for machines working 
at full pressure throughout the stroke, such as machine drills or 
simple, direct-acting pumps. When the motor works expan- 
sively the pulsations become objectionable, as a regular flow of 
air is not maintained in the return pipe. Under these con- 
ditions the inertia and friction of high-pressure air in long pipe 
lines becomes noticeable and disadvantageous. 

As the length of air pipe required for this system is doubled, 
not only may the first cost of the pipe go far toward offsetting the 
greater efficiency but, with at least twice as many joints in the 
pipe lines, the chances of loss from leakage are increased. And 
if very high pressures be used (pressures of several hundred 
pounds have been proposed), not only must the piping itself be 
heavier and more expensive, but the proportionate power loss 
from leakage is greater. For moderate distances, however, and 
when working at full pressure under the proper conditions, the 
foregoing disadvantages may be more than counterbalanced by 
the superior efficiency of the system. Though not yet in general 
use, the two-pipe system is said to have given satisfaction at sev- 
eral mines in New Mexico, Colorado, and California,* and has 
recently (1905) been proposed for use in the Johannesburg gold dis- 
trict. Some prominence is here given to the system because of 
its novel features and the probability that it may be found use- 
ful, if its disadvantages can be overcome. Reference may be 
made to a paper by H. C. Behr, published in 1905 in the Transac- 
tions oj the Mechanical Engineers^ Association oj the Witwaters- 
rand, in which the Cummings system is treated at length, with a 
discussion of its advantages as applied to compressed-air-driven 
pumps. 

* A. E. Chodzko, Modern Machinery (Chicago), Jan., 1899, p. 11. 



CHAPTER XVIII 

FREEZING OF MOISTURE DEPOSITED FROM COM- 
PRESSED AIR 

Reference has been made in a former chapter to the trouble 
sometimes caused by the congelation of the moisture carried 
in compressed air when deposited in the transmission pipes or in 
the ports and exhaust passages of the machine using the air. 
The presence of moisture in compressed air must be accepted as 
an unavoidable condition. Existing in the atmosphere at all 
times in greater or less quantity, when air is compressed the 
moisture is carried with it. A part of the water is deposited in 
the air receiver, but a considerable quantity still remains and will 
be brought into evidence when the proper conditions occur. 

The capacity of air for moisture depends primarily upon its 
temperature. Under ordinary atmospheric conditions i,ooo 
cubic feet of air contain about one pound of water. When its 
volume is reduced in the compressor cylinder, the increase of 
heat which takes place augments its moisture-carrying capacity. 
Any subsequent decrease in temperature reduces this capacity, 
and if the air be saturated the excess of moisture is deposited. 
Volume for volume, the capacity of air for moisture is independ- 
ent of its pressure or density. That is, at the same temperature, 
a cubic foot of air at atmospheric pressure will hold in suspension 
the same weight of water as a cubic foot at loo pounds pressure. 
But this must not be misunderstood. If a certain volume of 
moist atmospheric air be compressed isothermally, say to tV 
of its original volume, its water capacity is also reduced to tV) and 
To" of the water originally present in the air is deposited. There- 
fore, while the capacity for carrying moisture of a given vol- 
ume of air varies with the temperature, it must change also with 
any increase or decrease of pressure which changes its volume. 

274 



FREEZING or MOISTURE DEPOSITED FROM COMPRESSED AIR 275 

Causes of Freezing. Certain conditions are required to cause 
freezing of compressed air: deposited moisture must be present, 
and it must be subjected to a temperature below the freezing- 
point. So long as the temperature does not fall low enough, the 
presence of moisture can do no harm. Although one of the 
recognized functions of the air receiver is to permit the deposition 
of water before the air passes into the pipes, still, unless the re- 
ceiver be extremely large, the air leaves it warm — usually even 
quite hot — and therefore carries with it considerable moisture. 
In the case of wet compressors, unless liberal sprays are used to 
attain effective cooling, the air is apt to contain more moisture 
than that from dry compressors. A well-designed injection com- 
pressor, however, not too small for its work and therefore running 
at a moderate speed, will deliver cool air which will not give 
trouble from freezing. The air having attained nearly normal 
temperature before entering the pipe-line, its moisture-carrying 
capacity undergoes but little further reduction while passing 
through the pipe, and only a small amount of additional de- 
position takes place. With dry compression the percentage of 
humidity of the intake air, and the temperature at discharge, 
determine the quantity of water carried out of the cylinder. The 
humidity, in turn, varies with the weather. Changes in the 
weather may quickly be followed by variations in the quantity 
of moisture deposited in the receiver and pipe-line. When the 
air is finally expanded in doing its work in the air engine, intense 
cold is produced as the pressure falls, and the latent heat of 
compression is absorbed. It is here that the moisture carried 
with the air into the pipes makes its appearance as frost and 
causes trouble. Watery vapor itself, depositing a light, snow- 
like frost, does not tend to clog the air passages and ports as 
much as entrained water in a finely divided state, which will 
gradually form accumulations of solid ice and choke the exhaust 
wholly or in part. 

Prevention of Freezing. The difficulties which may arise 
from the conditions just outlined are apt to be exaggerated. That 
freezing not infrequently occurs is true, but with a properly 



276 COMPRESSED AIR PLANT 

designed and arranged plant it may easily be avoided. Two 
things require attention : prst, the air should be caused to drop its 
moisture as completely as possible before entering the main; 
second, provision should be made for draining off what deposited 
moisture remains in the pipe-line, before the air passes to the 
machine in which it is to be used. Although this is a simple 
matter, the means for accomplishing it are often neglected. Con- 
siderable quantities of water may collect in low places in the 
pipe-line and, if not blown out at intervals, will be carried into the 
ports, cylinder, and exhaust passage of the air machine and there 
freeze. 

Granting that the air leaves the receiver near the compressor 
practically saturated and still warm, it is evident that a great 
improvement in working conditions may be realized by intro- 
ducing a second receiver as close as possible to the machines 
using the air. In mining the second receiver is, of course, placed 
underground.* Before reaching it, the temperature of the air 
will have become normal, and the entrained moisture from the 
pipe-line may readily be trapped and drawn off. It may be re- 
marked that automatic water-traps are preferable to valves or 
cocks for getting rid of the water. As a rule, when the com- 
pressed air is to be used expansively, a special aftercooler should 
be introduced, placed as close as possible to the compressor. In 
any case, the receiver should be of ample size to insure the de- 
position of the moisture. The advantages of reheating the air 
before use will be taken up later. 

Influence upon Freezing of High Pressures in Transmission. 
The statements made in the first part of this chapter suggest an 
important consideration, viz: in transmitting power by air at a 
high pressure there is less liability to trouble from freezing than 
when low pressures are employed, provided that the length of 
pipe-line is sufficient to allow the air to be completely cooled and 
drained of its water while still under high pressure. At a low 
pressure a greater volume of air is required to furnish a given 
amount of power than when at a high pressure. More moisture ^ 

* See Chapter XI. 



11 



FREEZING or MOISTURE DEPOSITED FROM COMPRESSED AIR 277 

must, therefore, be dealt with, and at the low pressure it cannot 
be so thoroughly separated before the air is used. Suppose the 
transmission to be at a high pressure, and through a pipe long 
enough to allow the air to reach normal temperature. If the 
deposited moisture be drained away while the air is at its maxi- 
mum pressure; then, if the air be subsequently expanded down 
to a lower pressure suitable for working (with a corresponding 
increase of volume) and allowed to regain its normal tempera- 
ture, the percentage of moisture will be reduced, so that the air 
may be relatively very dry. When finally used in the air engine 
there will not be enough moisture present to cause troublesome 
freezing. 

Deposition of Moisture by Reducing Pressure. Still another 
mode of minimizing trouble from freezing is to reduce the press- 
ure of the air before it enters the cylinder of the air engine. The 
means by which this is accomplished and the results obtained 
may be illustrated by an example. 

At the Drummond Colliery, Nova Scotia, for running an 
underground pump by compressed air two receivers are used, one 
near the pump, and another 300 ft. farther back on the pipe- 
line. The air pressure in the main from the surface is 85 lbs., 
and as the proportions of the cylinders of this particular pump 
are such that so high a pressure was unnecessary a reducing valve 
was put in the pipe just before reaching the first receiver. By 
this valve the air is wire-drawn to reduce the pressure to forty- 
five pounds, which results in a deposition of nearly one-half the 
entrained water, in addition to that already deposited in the 
pipes. It is found that more moisture collects in the first than 
in the second receiver, and by this device the serious difficulty 
previously encountered from freezing at the pump has been en- 
tirely overcome.* The temperature lost by the reduction of 
pressure to forty-five pounds is regained before the air reaches 
the pump. 

* This information has been kindly furnished by Charles Fergie, superin- 
tendent of the Drummond Colliery. See also Mr. Fergie's article on the subject, in 
Transactions Canadian Mining Inst., 1896, of which an abstract was published in 
the Colliery Guardian, October 30th, 1896, p. 821. 



278 COMPRESSED AIR PLANT 

Protection of Surface Piping. What precedes refers only to 
the freezing produced by internal reduction of temperature, acting 
on the moisture carried in the air. In using compressed air, even 
for mining purposes, it often becomes necessary to carry lines of 
air pipe considerable distances on the surface. To prevent con- 
densation and freezing of the moisture in winter by external cold, 
all surface piping must be protected. If exposed to temperatures 
below the freezing-point, the inside of the pipe will become 
coated with ice and its effective cross-section reduced. A serious 
diminution of area may thus be caused at low points in the pipe- 
line, where water tends to collect ; or the pipe may even be frozen 
solid in such places by the gradual accumulation of ice. Under- 
ground the temperature is rarely, if ever, low enough to render 
any protection necessary, except in cold, down-cast shafts, or in 
tunnels in which there is a strong inward draught. 

Some time ago, at one of the Butte copper mines, a simple and 
inexpensive device was employed to prevent the freezing of mois- 
ture in a long line of surface piping. The air main of a large 
compressor plant was carried on the surface some hundreds of 
feet before reaching the shaft. During the winter months it was 
at times difficult to get sufficient air pressure in the mine because 
of the partial choking up of the pipe. As the volume of com- 
pressed air was too large to be dealt with by the ordinary receiver, 
a series of old tubular boilers were placed close to the compressor 
house. The hot air, at eighty pounds gauge pressure, in passing 
through these boilers, from one to another, was cooled down 
practically to atmospheric temperature and as a consequence a 
large part of its moisture was deposited. It was found that dis- 
carded tubular boilers, when strong enough, were well suited to 
this purpose, because of the large surface presented to the cold 
outside air; especially when they are set horizontally, so that there 
is a free circulation of air through the tubes. A blower might 
be used for the same purpose in a warm climate, or the boilers 
submerged in cold water. This effectual remedy is worthy of 
adoption where the conditions are similar. 



CHAPTER XIX 

REHEATING COMPRESSED AIR 

After the warm compressed air leaves the compressor and 
receiver on its journey through the transmission line its tem- 
perature is quickly reduced to that of the surrounding atmosphere. 
The loss thus suffered could be prevented only by using the air 
immediately and before it has time to cool. But this is never 
possible in mining practice. It would be unreasonable to pro- 
duce compressed air for use close to the compressor, because of the 
loss that inevitably ensues whenever power is converted from one 
form into another. The principal object in compressing air is 
to convert the power into a convenient form for transmission to a 
distance. The facility with which the heat of compression is 
given up, however, suggests that a gain may be effected by reheat- 
ing the compressed air when it reaches the place where it is to 
be utilized. 

The process is a simple one, and by such reheating an ad- 
ditional volume of air is obtained at a much lower power cost than 
if it were produced in the compressor itself. This may be shown 
by comparing the number of heat units required to produce a 
given volume of air at a given pressure in a compressor cylinder, 
with the heat units required to accomplish the same result by 
causing the air to expand through the direct application of heat. 
Herein is the ultimate basis of comparison for determining the 
useful effect of reheating. 

Appliances for, and Results of Reheating. A number of 
•methods of reheating have been actually used or proposed, the 
most important of which are as follows: (i) The air to be heated 
is passed through a cast-iron chamber or coil of pipe, exposed to a 
fire or current of hot gases or steam; (2) heat may be added 

279 



28o COMPRESSED AIR PLANT 

within the body of air itself, such as by the combustion of fuel, the 
injection of steam or hot water, or the placing in the air pipe of an 
electric-resistance coil. 

The method most frequently employed is the one first named; 
it is preferable from a mechanical standpoint and is the most 
efficient. Those appliances in which internal combustion is 
adopted, or in which hot water or steam is the heating agent, are 
less satisfactory in practical operation, but are useful where the 
burning of fuel is not admissible. 

The following calculation,* showing the results theoretically 
obtainable by reheating, presents the matter in concise form: 

Weight of I cu. ft. of steam, at 75 lbs. gauge = 0.2089 lb. 

Total heat units in i lb. of steam, at 75 lbs., produced from 
water at 60° F. = 1151. 

Total heat units in i cu. ft. of steam at 75 lbs. = 1151 X 
0.2089 = 240.44. 

To produce by compression in a steam-actuated air com- 
pressor I cu. ft. of compressed air at 75 lbs. gauge and 60° F., 
about 2 cu. ft. of steam at the same pressure are required,! 
making the thermal cost of i cu. ft. of compressed air, at the 
above temperature and pressure, 240.44X2=480.88 heat units. 
The air is here supposed to have lost its heat of compression by 
being stored or transmitted to a distance, so that the 480.44 heat 
units represent its cost at the motor where it is to be used. 

The result of reheating may now be stated : 

Weight of I cu. ft. of compressed air at 75 lbs. and 60° F.. 
= 0.456 lb. 

Units of heat required to double the volume of i lb. of free 
air at 60° F. = 123.84. 

Units of heat required to double the volume of i cu. ft. of 
compressed air at the same temperature and pressure = 123.84 X 
0.456 = 56.47. 

Comparing the thermal cost of i cu. ft. of air compressed 
in a cylinder with that of i cu. ft. obtained by reheating: 

* Frank Richards, " Compressed Air," p. 158. 

t That is, the efficiency of the compressor is assumed to be fifty per cent. 



REHEATING COMPRESSED AIR 28 1 

480.88 : 56.47 : : i : 0.1174 
that is, the cost in heat units of the volume of air produced by 
reheating is less than J of that required to produce the same 
volume by compression. 

It is not to be expected that the theoretical result here set 
forth can be attained in practice. To effect such a saving a very 
perfect form of reheater v^ould have to be employed, and the air 
after reheating pass directly into the cylinder of the engine. A 
farther conveyance of the air in pipes for even a very short dis- 
tance rapidly lowers its temperature and therefore its pressure. 

Temperatures Employed in Reheating. At a constant press- 
ure the volume of air is proportional to its absolute temperature, 
or 459° F. plus the sensible temperature above the zero point, as 
read on the thermometer. The absolute temperature of air at 
70° F. is 459 + 70 = 529°, In doubling the volume by the appli- 
cation of heat the absolute temperature becomes 1058°, and 1058 
— ^459 = 599°, which is the corresponding sensible or thermo- 
metric temperature. But this temperature is greatly reduced 
by the time the air reaches the motor cylinder, and still more 
heat is lost in the cylinder before its work is done. To reheat the 
air to a temperature which would really double its volume in the 
motor cylinder itself would involve a temperature in the reheater 
much higher than 599°. But such high temperatures cannot be 
employed, because they would render impossible the proper 
lubrication of the cylinder. If the temperature be raised by the 
reheater to 400° F. a loss of, say, 100° should be allowed for cooling 
before the air is actually used. The absolute cylinder tempera- 
ture is then 300 + 459 = 759°, and the corresponding added volume 
of compressed air practically available is found by the proportion: 

529 :759 ::i :i.43 + 
That is, there has been an effective increase of about 43 
per cent, in the volume of compressed air by heating in the 
reheater to 400°. It is improbable that a higher temperature 
would be desirable in the motor cylinder, or that any material 
further increase in economy could be realized in the operation of a 
compressed-air motor. In actual practice the gain derived from 



282 COMPRESSED AIR PLANT 

reheating is usually considerably less than is here shown. For 
air engines taking air at nearly full stroke, such as machine- 
drills and small, single-cylinder pumps, the increase of work 
ranges from, say, thirty to thirty-five per cent., without deducting 
the cost of the fuel used in the reheater. A higher efficiency is 
shown for expansive-working engines. 

For some purposes the determination as to the quantity of 
heat to be imparted in reheating is based on the temperature at 
which the air leaves the compressor cylinder, the idea being to 
recover the heat subsequently lost in cooling. Suppose, for ex- 
ample, that the compression is practically adiabatic, as is usually 
the case in single-stage dry compressors. Taking as the unit i 
lb. of air, or 13.2 cu. ft., at a temperature of 65° F., and com- 
pressing to 70 lbs. gauge, the heat of compression * is: 

= 869° absolute temperature, and the final thermometric tem- 
perature is, 869° — 459° = 410° F. The rise in temperature due 
to compression is therefore : 

4io°-65° = 345°F. 

If the compressed air be subsequently cooled to 65°, its volume 

14.7X1 3.2 

becomes: • — = 2.29 cu. ft. 

84.7 

In using this air without reheating and non- expansively, in a 

machine such as a rock-drill, having, say, 10 per cent, clearance, 

the work done is 

W = (2.29X144X84.7X0.9) — (2.29X144 X 14.7) = 20290 ft. lbs. 

But if the air be reheated to the final temperature of compression 

(345° F.), the work is: 

W'=: ^X 20290 = 33478 ft. lbs., and the work gained by 

reheating is therefore : 

33478 — 20290 = 13188 ft. lbs, or 39 per cent. 
The thermal cost of reheating this air will be: 345^X0.2375 

* See Chapter X. 



REHEATING COMPRESSED AIR 283 

(specific heat of air at constant pressure) =81.9 thermal units 

(B. T. U.), which are equivalent to 81.9X772=63226 ft. lbs. of 

work. 

Hence the efficiency of reheating in this case is : 

13 188 

= 20.8 per cent. 

63226 

In a series of experiments carried out in connection with the 
large plant of the Paris Compressed -Air Company, and using an 
improved form of reheater, the expenditure of coke in the heater, 
for one added horse-power per hour, was only 0.2 pound, which is 
say about one-eighth of the fuel consumption of large compressors 
of the best make, with compound steam cylinders. But with 
this particular plant the above very low fuel consumption in the 
heater was probably greatly exceeded. 

A working test, conducted by Prof. Alex. B. W. Kennedy, 
on a reheater supplying air for a small motor, gave the following 
results: The air was reheated to 315° F., with a consumption of 
about 0.39 lb. coke per indicated horse-power per hour, pro- 
ducing an increase of about 42 per cent, in the volume of the 
air, and, if the indicated efficiency had remained the same as 
during the trials with cold air, there should have been a decrease 

of air consumption in the ratio = 0.70. The volume of cold 

air used (admission temperature, 83° F.) was 890 cu. ft. per 
horse-power per hour; the volume when reheated was 665 cu. 
ft., or 75 per cent.; so that the full economy resulting from 
reheating was nearly realized. In this connection Professor 
Kennedy says : " I do not doubt that the stoking of the heater dur- 
ing my experiment was much more careful than it would be in 
ordinary practice, although, on the other hand, it would not be 
difficult to design a more economical stove. If, however, the coal 
consumption were even doubled, it would only amount to 72 
lbs. per day of 9 hours for 10 indicated horse-power, the value 
of which might be 6d. or 7d. The air saved per day under the 
same circumstances would be over 20,000 cu. ft., the cost of 
which, at the high rate charged in Paris, would be 7s. 3d." 



284 



COMPRESSED AIR PLANT 



A summary of the mean results obtained from two experiments 
on the above plant with cold, and two with reheated, air show:* 

1. With cold air. Incomplete expansion, wire-drawing, and 
other such causes, reduced the actual indicated horse-power of the 
motor from 0.50 to 0.39. 

2. By heating the air to about 320° F. the actual indicated 
horse-power at the motor was increased to 0.54. The ratio of 

0.54 



gain due to reheating was therefore 



0-39 



1.38. 



3. Deducting the value of about 0.39 lb. coke per indicated 

horse-power per hour, used in heating the air, the real indicated 

efficiency of the whole process becomes 0.47, instead of 0.54, and 

0.47 

the ratio of gain is reduced to ■ = 1.20=^. 

^ 0.39 

These carefully conducted experiments, though not made 

with a well-designed reheater, are valuable in proving that a sub- 
stantial net gain is obtained from reheating. Where reheating is 
employed in mine practice, how^ever, the quantity of heat im- 
parted to the air is usually much less than that indicated above. 
Good results may be obtained by the application of even less than 
100° F. 

The results of some experiments by Riedler and Gutermuth, 
on the consumption of reheated air, by an ordinary single-cylin- 
der eighty-horse-power engine, are given in Table XXVIII.f This 



Table XXVIII 



Test. 


Temperature of Air. 


Consumption 
Free Air per 
H.-P. Hr. in 
Cubic Feet. 


Indicated 
Horse-Power. 


Efficiency. 


Admission. 


Discharge. 


I 
2 
3 
4 


264.2° F. 
305-6 
320.0 
338.0 


69.8° F. 
84.4 
95-0 
120.2 


462.77 

431.09 

418.55 
432.12 


72-3 
72.3 
72-3 
65.0 


0.89 
.90 
.91 

.87 



* " Experiments upon the Transmission of Power by Compressed Air in Paris." 
Van Nostrand's Science Series, No. io6, p. 35. 
t Wm. Cawthorne Unwin, ibid., p. 104. 



REHEATING COMPRESSED AIR 



28s 



engine, with Corliss valve gear, was originally designed and used 
as a steam engine, and no changes were made for adapting it to 
work with compressed air, except that the cylinder was jacketed 
by the hot air on its way to the valve chest. The initial pressure 
was 95.5 lbs. absolute and the temperature of the air in the re- 
heater did not exceed 338° F., at a coke consumption of 0.176 
lb. per horse-power hour. 

Construction and Operation of Reheaters. The reheater em- 
ployed in the experiments referred to in the preceding section was 
that in use some years ago in connection with the Paris plant. 
It consisted of a double cylindrical box of cast-iron twenty-one 
inches diameter by thirty-three inches high, over all, enclosed in a 




Fig. 124. — Leyner Compressed Air Reheater. 

sheet -iron casing. The air under pressure traversed the annular 
space between the inner and outer cylinders, being compelled by 
baffle-plates to circulate in such a manner as to come into con- 
tact with the whole heating surface. The products of combustion, 
from a coke fire in the inner cylinder, passed downward over the 



286 



COMPRESSED AIR PLANT 




Fig. 125.— Cast-iron Coils, Ley- 
ner Reheater. 



exterior surface of the annular air chamber on their way to the 
chimney. A heater of this size served for a ten to twelve horse- 
power motor. 

In another form of reheater the air is passed through a coil 
of wrought-iron pipe, enclosed in a cylindrical casing. A large 
heating surface is thus obtained, but wrought-iron pipe is ob- 
jectionable because it burns out 
rapidly unless the fire is kept mod- 
erate. The conditions are materi- 
ally different from those to which 
boiler tubes are subjected, since the 
air tubing is denied the cooling 
effect of the water. Cast-iron coils, 
on the contrary, such as those of the 
Leyner reheater (Figs. 124 and 125), 
stand well. The U-shaped pipes 
are made in separate sections, bolted together as shown, with as- 
bestos packing in the joints. By varying the number of units 
any desired capacity can be obtained, and a broken or burned- 
out section is readily replaced. 

The Sergeant reheater (Fig. 126) consists of two concentric, 
cast-iron shells, bolted together, one within the other, the joints 
being packed with asbestos gaskets. The inner chamber forms 
the top of the fire-box. In shape this reheater is a truncated 
cone, set on a cylindrical fire-box, the cold -air main being con- 
nected by a flange coupling at the top and the hot air discharged 
near the base. This heater measures 42 ins. outside diameter 
at the base by 54 ins. high, with a grat^ ^19 ins. diameter. It 
is stated that 340 cu. ft. of free air per minute, at 40 lbs. pressure, 
can be heated to 360° F., with a gain of 30 to 35 per cent, in the 
energy developed. If more than 400 cu. ft. of free air per minute 
are to be reheated, 2 or more heaters of this size should be set in 
series, the air passing from one to another, allowing a maximum 
of 400 cu. ft. for each. 

The inner and outer shells of reheaters of the cast-iron- 
shell type are subjected to considerable differences of temper- 



REHEATING COMPRESSED AIR 



287 



ature, and except when of small size the upper and lower ring 
joints between the shells are difficult to keep tight.* In the 
Rand reheater (Fig. 127) the castings are more complicated in 
shape, the air passing between them in a thin sheet, from the in- 
let on the side to the discharge at the top of the central dome. To 
provide for expansion and contraction, the lower joint above the 




Fig. 126. — Sergeant Reheater. 

fire-box is provided with a stuffing ring and packing, shown in the 
cut. There is still a tendency to leakage, however, if the fire be 
very hot. 

The Sullivan reheater (Fig. 128) is quite different in design 
from those described above, consisting essentially of a vertical 
coil of cast-iron piping, or hollow rings, encased in double sheet- 
steel shells, the space between the latter being filled with asbestos. 

* Sibley Journal of Engineering, 1904. 



COMPRESSED AIR PLANT 



Below is the grate and combustion chamber, the gases from which 
pass through the spaces between the air rings. To minimize 
leakage, the centers of the rings are joined by malleable-iron 
nipples, so that all expand and contract together. These heaters, 




Fig. 127. — Rand Reheater. 



usually designed for burning coal, coke, or wood, are made in 
3 sizes, for 345 to 800 cu. ft. of free air per minute, having from 3 
to 7 rings, and measuring from 5 ft. 8 ins. to 7 ft. 6 ins. in height, 
by 33 ins. outside diameter. 

Internally fired reheaters — those in which the air is heated by 
direct contact with the fire — have hitherto been unsuccessful, 



REHEATING COMPRESSED AIR 289 

because dust and injurious products of combustion are carried 
by the air into the cylinder of the air motor. This trouble, of 
course, does not exist to the same extent when gasoline or other 
oils are used, instead of solid fuels, nor in the electric reheater, 
which, however, has thus far had 
but a limited application. 

A fault of most reheaters as 
built at present is that there is no 
provision for regulating the heat ac- 
cording to the variation in con- 
sumption of air, such as is un- 
avoidable in applying reheating for 
machine drills, channellers in 
quarry work, hoisting engines, and 
other intermittently operating ma- 
chinery. This want of regulation 
evidently is not so important for 
constant-running engines, such as 
pumps. 

As the air chamber, of what- 
ever shape, in all of the externally 
heated or'' dry" reheaters, forms in 
reality a part of the air main, 
reheating can increase the press- 
ure only in a small degree. Its 
real effect is to increase the volume 
of air, which tends to back up 
in the main, reducing incidentally 
the velocity of flow and there- 
fore the loss of pressure due to friction. The reheater should 
always be placed as close as possible to the machine using the air. 
This is readily done with stationary engines, like pumps or 
hoisting engines; and even in the case of movable machines, like 
quarry channellers, the reheater may be set on the same carriage 
or bed-frame. If it be necessary, however, to convey the heated 
air some distance, the temperature may be quite effectually 




Fig. 128. — Sullivan Reheater. 



1 



I 



290 COMPRESSED AIR PLANT 

maintained by covering the pipe with non-conducting material, 
as is done with steam piping. 

Sometimes when the air-engine cylinders are compounded, 
the exhaust from the high-pressure cylinder is passed through 
a second reheater before going to the low-pressure cylinder. A 
further benefit may be derived by injecting into the reheater a 
very small quantity of water. The specific heat of water is high 
as compared with the specific heat of air; also such part as is 
converted into steam gives up its latent heat in the motor-engine 
cylinder and prevents trouble from freezing, even when a high 
rate of expansion is employed. For the same reason, benefit 
may be derived from injecting a little warin, or even cold, water 
into the compressed-air feed-pipe of an air motor. Water used 
in this way acts incidentally as a mechanical scourer, in washing 
away accumulations of ice tending to form in the ports. 

It will be seen from the construction of reh eaters that the 
calorific power of the fuel burned in them is not economically 
utilized. The flue loss is large for the same reasons that apply 
to the work of ordinary shell or tubular boilers. But the thermo- 
dynamic advantage gained is so considerable that the low 
efficiency of the reheater itself, in burning the small quantity 
of fuel required, becomes of secondary importance. 

Use of Reheaters for Underground Work. In the ordinary 
operations of mining the reheating of compressed air can have 
only a limited application. By far the most important use of 
compressed air in mining is for operating machine drills. Up to 
the present time there are relatively few mines where it is em- 
ployed for any other purpose. But it is evident that for portable 
machines like rock-drills, continually being shifted from place to 
place underground, the use of reheaters in most cases is economic- 
ally out of the question. Not only would a number of them be 
necessary, but they would have to be moved about and kept close- 
to the drills, to prevent the reheated air from losing its heat and 
temporary increase of volume. 

For stationary engines, however, such as underground pumps, 
hoists, rope-haulage engines, etc., and wherever the reheater can 



REHEATING COMPRESSED AIR 291 

be placed permanently close to the air engine, reheating in mines 
may be successfully applied. The idea that it is useful mainly 
in preventing the accumulation of ice in the exhaust ports and 
passages of the air engine is not an uncommon one; but as a 
matter of fact the prevention of freezing is merely incidental to a 
decided gain in efficiency. In underground work it may be 
difficult to arrange for burning fuel under a reheater, notwith- 
standing the small quantity required, because of the resulting 
vitiation of the mine atmosphere. Also, in gassy collieries re- 
heaters, cannot well be used, though sometimes the products of 
combustion may be turned into an upcast airway, or even allowed 
to escape into the mine workings, when the heater is small and the 
active circulation of large volumes of air is maintained. Where 
the conditions underground are such that strong combustion is 
not allowable and only a small quantity of fuel can be burned in 
the reheater, it will still be found that some advantage is obtain- 
able from air engines by a very slight added temperature — say, 
only 25° to 50° F. In this connection it may be noted that the 
use of the internal electric reheater, already referred to, in which a 
resistance coil is placed in an enlarged section of the air main^ 
does away with the difficulty of disposing of the products of com- 
bustion of fuel and would be especially useful in gassy collieries. 
Another mode of applying electric reheating is to wrap the resist- 
ance coils around a short length of the air pipe. 

At the North Star Mine, Grass Valley, Cal., the plan has been 
adopted of placing a reheater on the surface near the shaft mouth 
and carrying the compressed air underground by a pipe covered 
with non-conducting material. Fairly satisfactory results are thus 
obtainable, with the advantage of avoiding the burning of fuel 
in the mine. But while some saving can be realized in this way 
for moderate distances — say of a few hundred feet — it would be 
economically out of the question for long transmission lines. 
This arrangement suggests the caution that non-conducting 
covering should always be used for the pipe from reheater to air 
engine, however short the distance. In a case on record,* where 

* Richards, American Machinist, Feb. 28th, 1895. 



292 COMPRESSED AIR PLANT 

this distance was only 20 ft., but no pipe covering provided, the 
gain in power realized was only 12 per cent., though the absolute 
temperature of the air was increased at the reh eater 38 per cent., 
with of course the same theoretical increase of volume. The 
air used for operating an underground pump at another Cali- 
fornia mine is reheated by steam conveyed from the surface. 
Steam may thus be used to greater advantage than if employed 
directly in the cylinder of a pump; for, in condensing, the la- 
tent heat otherwise lost is utilized in raising the temperature of 
the air and is so converted into work. All devices of this kind, 
however, must be classed as makeshifts. 

In recent years several mine plants have been erected at which 
compressed air has been used even for operating surface hoisting 
engines — for example, at the Lightner Mine, Calaveras Co., 
Cal. One of the units of a battery of boilers is adapted as a 
reheater. The compressed air passes from the receiver into a 
section of perforated pipe submerged just below the surface of the 
hot water in the boiler, and is thence led to the hoisting engine. 
By means of a large receiver capacity, quite satisfactory results 
are secured, notwithstanding the intermittent work of the engine. 

In connection with the method of reheating referred to above 
the results may be given of a number of experiments made by 
Prof. J. T. Nicholson, in reheating air from the Taylor Hy- 
draulic Air Compressor, at Magog, Prov. Quebec, described in 
Chapter XV. The air was used in a 27-horse-power Corliss 
engine, at a pressure of 53 lbs. There were 5 tests, as follows: 
I. With cold air. 2. Reheating by means of steam injected into 
the air main before reaching the engine. 3. The compressed air 
was passed into a steam boiler, and heated by mixing with the 
water and steam. 4. The compressed air was blown upon the sur- 
face of the water in the boiler, and heated by mixing with the steam. 
5. The air was passed through a tubular reheater, fired by coke. 

Without reheating, 850 cu. ft. of free air were used per in- 
dicated horse-power hour. By reheating in the boiler, a mix- 
ture of 10 to 15 lbs. of steam with the air reduced the consump- 
tion of air from 850 cu. ft. to 300 to 500 cu. ft., per indicated 



REHEATING COMPRESSED AIR 293 

horse-power hour. Thus, i added horse-power was obtained 
by wet heating, at an expenditure of from i to 1.3 lbs. of coal 
per horse-power hour. 

By dry heating in the coke-fired reheater, the air was raised 
to 287° F. At this temperature, 640 cu. ft. of free air were re- 
quired per horse-power hour, or 210 cu. ft. less than with cold air, 
the saving in the quantity of air being about 25 per cent. By 
burning in the reheater 47 lbs. coke per hour, 100 horse- 
power in cold compressed air was raised to 133 horse-power, 
making-an expenditure of 1.42 lbs. coke per hour for each added 
horse-power. These results indicate that the reheater used was 
not very efficient. But though the fuel consumption was much 
greater than in Professor Kennedy's test, previously described, 
it is still far lower than is attainable in the most efficient engine 
and boiler practice. 

In a paper by Clarence R. Weymouth, on "Reheating Com- 
pressed Air with Steam,"* a detailed discussion is given of the 
thermodynamics of this mode of procedure, with deductions as 
to its efficiency. The author considers the cases of injecting 
steam into the air main, and of passing the compressed air through 
a steam boiler, giving the results in tabulated form. 

* " Compressed Air Information." Edited by W. L. Saunders. 



CHAPTER XX 

COMPRESSED AIR ROCK DRILLS 

In the introductory chapter reference was made to a few of 
the facts relating to the earlier stages of development of the modern 
rock drills, and to the importance of these machines in the working 
of mines and quarries, the sinking of shafts and the driving of 
tunnels. The machine drill has not only been the means of in- 
creasing greatly the speed of work, thereby reducing the cost of 
all operations involving rock excavation, but it has made possible 
the driving of long tunnels, which, it is safe to say, could never 
have been completed by hand drilling. 

It is unnecessary here to trace the history of machine drills. 
This has been well done in Drinker's treatise on '' Tunneling, 
Explosive Compounds, and Rock Drills," first published in 1878. 
In that book details are given of many of the numerous patents 
taken out in the United States and Europe, from 1849 to 1882, 
and the successive steps in the earlier development of the machine 
drill are fully recorded. The present chapter will be devoted 
chiefly to the description of a number of representative drills now 
in use, together with notes on the performance, consumption of 
air and other matters relative to the operation of machine drills. 
While it may not be said that rock drills have become so standard- 
ized in design that further important improvements are improbable, 
yet, in the past eighteen or twenty years, and especially since the 
"air hammer" drill was introduced, radical changes in principle 
have been few. Weak points in design and construction have 
been discovered by experience, and remedied as far as practicable, 
so that at the present time there are a number of successful, 
serviceable machines on the market. 

For rock work the percussion drill only is of practical use, at 
least for any rock harder than soft shales, coal, and other similar 

294 



COMPRESSED AIR ROCK DRILLS 295 

material. All attempts to construct a rotary pneumatic rock drill 
have thus far failed. The diamond, and other core, drills for deep 
boring, and the efficient Brandt rotary drill, operated by hydraulic 
power, obviously have no place in the present discussion. 

General Description. The reciprocating or percussion rock 
drill, aside from those machines that operate on the hammer 
principle (see Chap. XXI.), may be roughly described as consist- 
ing of a cylinder, in which either compressed air or steam is used, 
the drill bit being firmly clamped to a chuck, forming the end of 
the piston rod (Fig. 129.) For admitting and exhausting the 
compressed air alternately at each end of the cylinder a number 
of different valve motions are in use. Some of these are similar 
to the valve motions of certain single-cylinder pumps; in others 
a positive movement of the valve is caused by the introduction of 
a tappet, actuated by the strokes of the piston. There is also a 
device to produce automatically the necessary rotation of the drill 
bit on its axis, for keeping the hole round and preventing the bit 
from sticking. In standard drills this is done by a rifle bar, ratchet 
and pawls. Working speeds are usually from 300 to 400 strokes 
per minute, for the larger sizes of drills, up to 500 strokes for the 
smaller; the normal length of stroke, in drills of average size, being 
from 4J to 6 inches. The admission of air, and therefore the speed 
and force of the blow delivered, are controlled by a hand valve in 
the air pipe close to the valve chest. 

A feed screw, with crank and handle, is carried in a bearing 
at the rear end of the shell supporting the cylinder, and engages 
with a nut on the under side of the cylinder casting. By this means, 
the entire drill head is fed forward by hand as the hole is deepened, 
several bits of successively greater and greater length being clamped 
as required in the piston rod chuck. (An automatic feed has been 
introduced, and used to some extent for surface work, but is 
neither necessary nor entirely satisfactory for underground service.) 
When the cylinder has been fed forward as far as the length of the 
screw and of the shell will permit, the drill is stopped. By revers- 
ing the feed the cylinder is moved back on its supporting shell, 
the bit removed and a longer bit put in the chuck. The cylinder 



296 COMPRESSED AIR PLANT 

is then fed forward until the new bit touches the bottom of the 
hole, with the piston nearly at mid-stroke; the air is turned on 
slowly and the work proceeds. It will be seen from this that the 
length of stroke, and therefore the force of the blow, are under the 
control of the drill-runner. If, while the machine is at work, 
the cylinder is fed forward faster than the hole is being deepened 
the stroke necessarily becomes shorter and shorter, because the bit 
strikes the bottom of the hole before the full length of stroke 
is reached; on the other hand, should the feed be too slow, the 
piston will strike the front cylinder head. Thus, the force of the 
blow may be graduated to suit the conditions. When starting a 
hole, for example, the stroke should be shorter than when some 
depth has been reached and the bit has adjusted itself to the shape 
of the bottom of the hole. Moreover, for hard rock, a short, 
rapid stroke gives the best results; while a longer stroke may be 
adopted for softer or tough rock. 

Further details of the construction and operation of machine 
drills are given hereafter, in describing examples of the different 
makes. 

Modes of Mounting Drills. The drill head, comprising the 
cylinder and its appurtenances and the supporting shell, may be 
mounted either on a tripod or column. For surface work the 
tripod only can be used; underground, either the tripod or column, 
depending on the size and shape of the working in which the drill- 
ing is to be done. 

I. Tripod. (Fig. 129.) The legs, which are telescopic, are 
hinged by a heavy bolt to the tripod head, and can thus be set 
as necessary for adjusting the position of the axis of the cylinder 
and bit, for the hole to be drilled. To the tripod head is bolted 
the "shell," which is provided with guides for supporting the 
cylinder, as it is fed forward. After the machine has been placed 
in position for drilling all bolts are tightened up. Heavy weights 
are usually slipped on the tripod legs, to prevent the drill from 
shifting while in operation. 

Tripod mountings are required not only for surface drilling, 
but also for underground work, when the distance between roof 




Fig. 129. — Sullivan Tappet Drill. 



298 



COMPRESSED AIR PLANT 



and floor, or between the side walls of the workings, is too great 
to permit the use of columns. Where a choice exists, however, 
the tripod is sometimes employed in preference to the column; 
because, as a rule, it may be set up with less loss of time and allows 




Fig. 130. — Double-Screw Column Mounting for Rock Drills. 



greater freedom in locating the holes, to produce the best results 
under existing conditions as to character of rock and shape of the 
face. The same may be true, also, in sinking shafts of large cross- 
sectional area, or when the rock is so irregularly fissured that a. 
symmetrical arrangement of the holes is undesirable. 



COMPRESSED AIR ROCK DRILLS 299 

2. Column or Bar, (Fig. 130.) The cut shows the usual form; 
a hollow, iron bar, varying in diameter from 3 in. to 5 J in., accord- 
ing to the size and weight of the drill. It can be used only under- 
ground and is securely set between the roof and floor, or the walls, 
of the working. The lower end of the column has a cross-piece 
or base, through which pass a pair of jack-screws. The upper 
end terminates in a serrated cap or head, which, by tightening up 
the screws, takes a firm hold upon the surface against which it is 
pressed. i\nother form of column has a single, telescopic jack- 
screw;- it may be used in small tunnels or mine workings, and also 
for shaft-sinking. In mine workings of large size the double- 
screw column is preferable, the drill being carried on an arm, 
attached to a collar encircling the column. When necessary two 
drills are mounted on the same column. The collar, or collar 
and arm, slides on the column longitudinally, and may also 
be revolved around the column. It is clamped fast in any posi- 
tion, as desired, for adjusting the height and angular direction 
of the drill. For the single-screw column the drill is attached 
directly to the collar, the arm being omitted. 

The column mounting is specially useful in driving mine 
tunnels, cross-cuts, and drifts, and for the advance headings of 
railroad tunnels. It is frequently employed, also, for stoping, 
when the pitch of the vein and the method of mining make it in- 
convenient to use tripods. When placed in an approximately 
horizontal position, as in shaft-sinking, the column is known as a 
"bar," though the mode of mounting the drill upon it is sub- 
stantially the same. Shaft-bars are sometimes made extra long, 
for wide shafts, in which case a pair of adjustable legs are hinged 
to a collar at the middle point to carry the weight of the drill or 
drills without making the bar inconveniently heavy. 

Formerly, for driving railroad tunnel headings, four, six or 
more, machine drills were mounted on a carriage, running on rails 
like a car, but these are no longer used in general practice. 

Classification of Compressed Air Drills. There are two distinct 
classes : First, Reciprocating drills, a name which may be given to 
those in which the drill bit is firmly attached to the piston rod, and 



300 COMPRESSED AIR PLANT 

delivers a succession of blows on the bottom of the hole; second, 
Air Hammer drills, in which the bit does not reciprocate, but is 
held in the forward end of the cylinder, and is struck by the piston 
as by a hammer.* 

RECIPROCATING DRILLS. 

These constitute the more important of the two classes and 
comprise a number of types and makes. Among the best known 
American drills are the Doble, Ingersoll, Jeffrey, McKiernan, 
Murphy, Rand, Sergeant, Sierra (Rix), Sullivan, Temple "Air- 
Electric," and Wood, Some of these are widely used throughout 
the world. Of the European reciprocating drills the following 
may be mentioned: Adelaide, Barrow, Chersen, Climax, Darling- 
ton, Ferroux, Froelich, Hirnant, Holman, "Konomax," Kiizel, 
Little Wonder, Meyer, Rio-Tinto, Schram, Triumph, and 
''Wahrwolf." 

Based on the design of the valve-motion, reciprocating drills 
may be further classified as: tappet-valve and spool- or piston- 
valve machines, and those in which the valve is eliminated com- 
pletely, its function of controlling the admission and exhaust of 
the compressed air being exercised by the piston itself. Examples 
of each form are given in the following pages. 

" Sergeant " Rock Drill. Fig. 131 shows a longitudinal section 
of this machine, which is built by the Ingersoll-Rand Company. 
The spool-valve and the main air and exhaust ports are similar in 
some respects to those of a single-cylinder pump. Air is admitted 
on one side of the valve chest, the exhaust opening being on the 
other side. 

The valve-motion is non-positive and consists of two parts: 
a spool-valve, which controls the main cylinder ports and an arc- 
shaped tappet, set in a correspondingly curved slot or seat, as 
shown, between the cylinder and valve chest. The ends of this 
tappet, which project slightly into the main cylinder, are struck 
alternately by the front and back shoulders of the large annular 

* Air Hammer drills are discussed in Chapter XXI. 



302 COMPRESSED AIR PLANT 

recess in the middle of the piston, thus causing the tappet to 
oscillate at each stroke of the piston. Behind the tappet, and in 
the vertical wall of its seat, are three small auxiliary ports, one in 
the middle and one near each end of the seat. These auxiliary 
ports connect with the spool- valve chest above; the middle port 
with the middle of the chest, the rear port {i.e., nearest the back 
end of the cylinder) , with the forward end of the chest and the 
forward port with the rear end of the chest. In the face of the 
tappet is a curved slot, just long enough, when in the extreme 
positions of its throw, to form a communication between the middle 
auxiliary port and one of the end auxiliary ports. That is, at each 
stroke of the main piston, the tappet places the opposite end of the 
valve chest in communication with the exhaust, thus causing the 
throw of the spool-valve. In the peripheral surface of the spool- 
valve a very fine longitudinal slot is cut, which constantly admits 
a small quantity of live air to both ends of the chest. Hence, when 
either end of the chest is connected with the exhaust, as stated 
above, the valve is thrown towards that end by the air pressure in 
the other end of the chest. 

In Fig. 131, the piston is beginning its forward stroke. The 
spool-valve, in its rear position, is admitting air to the back end 
of the cylinder, while the forward end of the cylinder is connected 
with the exhaust. As the piston advances, the rear shoulder of 
the annular recess in the piston strikes the projecting end of the 
tappet and throws it over. This changes the relation between the 
auxiliary ports, already described, exhausts the air from the front 
end of the chest and throws the spool-valve forward, thus prepar- 
ing for the back stroke of the piston. 

By the introduction of the arc-tappet, the Sergeant drill avoids 
in large part one of the chief defects of the ordinary spool-valve 
drills, viz.: irregularities in the operation of the machine, caused 
by wear of the valve and seat, which permits leakage of the air or 
steam past the valve. Adjustments for any irregularities of stroke 
produced by wear are made by the simple compensating device 
shown in Fig. 132, which is an enlarged section of the valve and 
chest. A hollow brass plug, P, having a very small hole, Hy 



COMPRESSED AIR ROCK DRILLS 



303 



permits the passage of a little live air to the back end of the chest. 
Should the piston strike the back cylinder head, the area of H is 
reduced slightly by peening or riveting up the edge of the hole. 
This decreases the quantity of air passing to the end of the chest 
and increases the cushioning in the rear end of the cylinder. If 
the stroke be too short, H may be found partly obstructed and 
should be cleaned, to admit more air to the end of the chest; if 
the stroke be still too short, the area of H is slightly increased with 
the point of a knife blade. 

The rotation of the piston and bit is caused as follows: A 
rifled bar, with a ratchet head and pawls set in the rear cylinder 



Front End 



B^iCKEnd 




Fig. 132. — Spool- Valve and Chest. "Sergeant" Rock Drill. 



head, engages with a correspondingly rifled nut screwed into the 
end of the hollow piston. The teeth of the ratchet wheel and its 
pawls are so shaped that, on the forward stroke, the piston moves 
without rotation, the rifle-bar turning the ratchet in its seat. On 
the back stroke the pawls prevent the ratchet from turning, so that 
the piston is compelled, by the rifle-bar and nut, to rotate through 
a part of a revolution. The ratchet ring, with internal teeth, with 
which the pawls engage, is not fastened rigidly in the back cylinder 
head, but is held by friction only, under pressure of an exterior 
cushion spring, acting on the periphery of the ratchet. Hence, 
when the drill bit sticks in the hole, or for any reason cannot rotate 
freely on the back stroke, the ratchet itself turns, thus preventing 
injury. The drill is fed forward on its supporting shell by a long 



304 



COMPRESSED AIR PLANT 



feed screw, engaging with a feed nut in a lug on the under side of 
the cylinder. 

The Sergeant drill is built in seven sizes, the cylinders measur- 
ing: 2 in., 2I in., 2 J in., 2f in., 3 in., 3} in., and 3 J in. diameter; 
weights of drill-head, unmounted, range from no to 405 lbs. 

Sullivan " Differential " Drill. Fig. 133 shows this drill as 
designed specially for steam. While the design and construction 
are essentially the same, the spool-valve of the air drill is ground 
with a larger clearance, to reduce the danger of freezing when 
the air is exhausted. Also, the front head is modified. Instead 
of the soft, steam packing 29, and the gland 28, a metal lining is 
used, with a leather packing ring. Several types of air head 
are designed for this machine, a bushing being sometimes pro- 
vided. Fig. 134 shows this drill as designed for using com- 
pressed air. 

Referring to F:g. 133, the chest 2 contains the spool-valve 6, 
air inlet port 3, exhaust ports 4, 4, and cylinder ports 35, 35. The 
valve is thrown by the action of the small reverse ports 5, 5, com- 
municating from the ends of the chest to the opposite ends of the 
cylinder as indicated. In the cut, the piston is shown on its 
forward stroke. When its rear end passes the opening into the 
cylinder of the right-hand reverse port 5, live air is admitted to 
the opposite, or left-hand, end of the valve chest; thus throwing 
the valve to the right, to prepare for the back stroke of the piston. 
During the movements of the valve, the cylinder ports 35 are 
alternately placed in communication with the exhaust ports 4, 4 
and the air inlet 3. 

Rotation of the piston and bit, on the back stroke, is produced 
in the usual way by the rifle-bar 13 and the ratchet head 14. The 
ratchet has broad teeth with rounded surfaces, and steel pins or 
rollers are provided, instead of ordinary pawls. To prevent 
injury, in case the bit should become wedged in the hole and so 
resist rotation, the ratchet ring, as usual, is held in its seat by 
friction only. In addition to the piston rings 34, the piston carries 
two collars 36, provided with longitudinal slots, for obtaining 
efficient circulation of the lubricating oil. Oil is admitted from 




M 



s 

- a; 
.2 



f5 S . -g 2 

C^_OgooOCO 

^ rt .^2 iS i .S .2 f, .^ 



c 

?! 
O 

o 2 



_g -a =■ c ^ 6 ^ 

T^ o) (11 3 w, Jj tn 



'- - H fe o m 



►^ 53 ^'j ,<u ^ 



•ri 


d 6 ^ ^ ^ 


Tf 


If) O r^ 




M N M Pt « 


H 


O C) PI 


<u 
















OJ 








fC 
















Q 








c 








c3 








> 


















ul M 




h 


1 


• lU C 




<u 


CO 
CO 


ckhead 
chamb 
Ltchet ri 
fie bar. 
Ltchet. 


2? 


ed nut. 
ck wash 
eck nut 


H 


pqOP^^fi^Sfe^o 


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\^ 


O H (N fO ^ 


l/^ 


O r- 00 



.■ai 



>, ,i^ '^ X <ii ^ '^ :2 ^ 



P) CO ■^ 1/^ vo 




Oh 



3o8 



COMPRESSED AIR PLANT 




COMPRESSED AIR ROCK DRILLS 309 

the chamber ii, through a small passage, to the ratchet-head and 
thence to the cylinder itself. 

The "differential" (spool-valve) drill is made in ten sizes, with 
cylinder diameters of: 2 in., 2 J in., 2 J in., 2| in., 3 in., ^l in., 
3} in., 3f in., 4J in., and 5 in. Weight, unmounted, from no to 
680 lbs. 

Sullivan Tappet Drill. (Figs. 135 and 136.) In its general 
lines this is similar to the spool-valve drill of the same makers. 
Air is admitted (Fig. 136) at the connection 6 to the chest 5, 
which contains the slide valve 4, controlling the air and exhaust 
ports. The tool-steel rocker or tappet 3 is not pivoted, but 
oscillates in an arc-shaped slot, as it is struck alternately by the 
two beveled shoulders in the middle part of the piston. On the 
back of the rocker is a lug, of standard rack-tooth form, which 
engages with a corresponding socket in the underside of the valve 
4, thus throwing the latter. Between the cylinder casting and 
the valve-chest is the valve plate 2, which may readily be re- 
moved for obtaining access to the rocker. 

The rotation and feed mechanism are the same as in the 
'' differential" drill. The band 9, around the front cylinder head 
7, is provided with lugs for the side bolts, which run back to the 
rear head. On each bolt, forward of the band, is a heavy spiral 
spring, to take up the shock in case the piston should strike the 
head. In the cylinder head is a leather packing-ring 8. 

The Sullivan tappet drill is made in four sizes, whose cylinder 
diameters are: 2| in., 3I- in;, 3^ in., and 3I in.; weights, un- 
mounted, from 233 to 393 lbs. 

Jeffrey " Badger " Drill. Originally, a machine called the 
"Badger" was made by the Phillips Rock Drill Co. In an im- 
proved form it is now built by the Jeffrey Manufacturing Co. 
Fig. 137 is a general longitudinal section. Though of the spool- 
valve type, it differs materially in some of the features of its valve 
motion from the drills of the same class already described. 

The valve a is a double, balanced spool, whose front end is 
of larger diameter than the rear. The air inlet is on the side of 
the machine, just below the chest, the exhaust being on the up- 



;io 



COMPRESSED AIR PLANT 



per side of the chest itself. 
Fig. 138 is a diagram show- 
ing, in a necessarily distorted 
manner, the relations and 
connections of the system 
of main and auxiliary ports, 
which cannot be completely 
represented in the longitudi- 
nal section. At the beginning 
of the outward stroke, air 
enters through the port h, to 
the annular recess in the 
main piston; passing thence, 
by the lower port c, to the 
rear end of the cylinder. 
Simultaneously, the forward 
end of the cylinder and the 
rear end of the valve chest 
are exhausting respectively 
through the main port d^ 
and the auxiliary ports, e 
and/. The valve is held in 
the position shown by the air 
pressure in the forward end 
of the chest, previously ad- 
mitted through the auxiliary 
port g. When the piston 
has advanced a short dis- 
tance, it closes the port c 
and opens the upper main 
port h. In this position of 
the piston air is admitted to 
the rear end of the cylinder 
through ports h and i to the 
valve, and thence, as indi- 
cated by the arrows, through 




COMPRESSED AIR ROCK DRILLS 



311 



port h. Advancing still farther, the piston closes communication 
between the air inlet and the rear of the cylinder, by covering the 
port i, so that the stroke is finished by expansion of the air al- 
ready admitted. 

At this point, the annular recess of the piston begins to uncover 
the auxiliary port /, which admits air to the rear end of the valve 
chest. The valve now reverses by the difference of pressure of 
the live air on the small end, and of the expanded air from the 
cylinder (through port g) on the forward or large end, of the 
valve. This reversal prepares for the back stroke of the piston 




Fig. 138. — Jeffrey "Badger" Drill. Diagram of Valve and Ports. 



by allowing air to enter by the auxiliary port /, to the chest and 
thence through the main port d, to the forward end of the cylinder. 
Toward the end of the back stroke, the rear ports h and k are 
closed by the piston and the back stroke is thus cushioned. The 
forward stroke is uncushioned. 

The Badger drill is made in four sizes, the diameters of cylinder 
being: 2| in., 2| in., 3I in., and 3! in.; maximum length of stroke, 
from 5 in. to 7 J in. Weight of drill, unmounted, ranges from 150 
to 320 lbs. 

Ingersoll-Rand " Arc-Valve " Tappet-Drill. (Fig. 139.) This 
machine furnishes an example of a positive valve-motion and 



312 COMPRESSED AIR PLANT 

constitutes an improvement on the earlier tappet drills of the same 
makers, which have a straight-face slide valve. 

The valve motion is as follows: A three-armed rocking 
tappet is pivoted in a recess in the upper part of the cylinder. 
When the drill is at work, the ends of the two lower arms of the 
tappet are struck alternately by the front and back shoulders of 
a wide, annular recess in the piston, thus causing the tappet to 
oscillate on its fixed center. The third arm, projecting upward, 
engages with a socket in the back of the slide-valve, and throws 
the valve positively, in an arc-shaped path, at each stroke of the 
piston. The seating surface of the valve, instead of being plane, 
is therefore also circular, its center being the center of the tappet 
pin. No auxiliary ports are required in this simple design; the 
valve controls the two main cylinder ports and the exhaust, as 
shown. In addition to the usual cushioning on the back stroke, 
common in all rock drills, the forward stroke is also slightly 
cushioned; this result being produced by the shape of the lower 
arms of the tappet, the recess in the piston and position of the 
main ports. The machine is provided with the usual feed and 
rotation mechanism. 

The drill may be operated by either compressed air or steam. 
It is found to be better adapted to steam than the piston valve 
drills, which do not work satisfactorily with wet steam, or in the 
presence of water of condensation. 

The " Arc- Valve " tappet drill is made in six sizes, the cylinders 
measuring 2J in., 2J in., 2| in., 3J in., 3J in., and 3! in. diameter; 
weights of the drill head, unmounted, are from 140 to 415 lbs. 

Murphy " Little Champion ** Drill. Fig. 140 shows the general 
plan, and longitudinal and cross-sections, of this machine, which 
has been on the market about ten years. The rotation mechanism 
is similar to that of other drills already described, and is clearly 
indicated in the sections. The valve-motion is of the tappet type, 
the upper arm of the 3-arm tappet a engaging with the flat slide 
valve h. This valve controls the main ports c, c and the exhaust 
port, which is on one side, opposite the middle of the tappet. A 
novel feature of the drill is the mode of producing the tappet's 




p 

a; 

Ph 



f^ 



314 



COMPRESSED AIR PLANT 



© 



© 



© 



P 



© 





SECTION SHOWING RATCHET 

WHEEL, PAWLS, PLUNGERS, SPRINGS 

AND PLUNGER GUIDES 



Weight of Drill = 280 Lbs. 

Extreme Length = 46 Inches 

Extreme Height = 12>4 " 

Width of Drill =.-== 9H 

Extreme Width of Drill 1 - ^q 
Including Air Tap 

Length of Feed ^-= 26 

Strolies per MMaute — From 450 

to 500 
According to 
Air Pressure 




-inch. 



PLAN OF CRADLE, SHOWING ONE SLIDE AND ALL BOLTS REMOVED 



COMPRESSED AIR ROCK DRILLS 315 

movements. Its two lower arms do not project into the top of 
the cylinder bore, to a contact with the piston, as is the case with 
most drills of this type. Instead, a steel pin d, sliding in a bushed 
hole or seat in the cylinder casting, is set under each tappet arm. 
As the piston makes its strokes, the curved shoulders of the annular 
groove c, in the piston, alternately strike the rounded ends of the 
forward and rear tappet pins, d, d, thus pushing up the pins and 
causing the tappet to oscillate. 

The Murphy drill is made in eight sizes: of 2 J, 2 J, 2f, 3, 3I, 
Sh Sh ^^^ 3f i^ch diameter of cylinder, the drill head and shell 
weighing, unmounted, from 125 to 395 lbs. It is designed for the 
usual tripod or column mounting. 

Climax Imperial Drill. (Fig. 141.) This is a well-known 
English machine, of the spool- or piston-valve type, made by 
R. Stephens & Son, Carn Brea, Cornwall. Air enters the valve- 
chest by the air tap, shown in detail section and also in place on 
the main elevation. Thence it passes into the annular recess b, 
of the valve (main longitudinal section), which, by its reciproca- 
tions, opens communication through the valve-seat ports c, alter- 
nately with the main cylinder ports a, a. In the valve are the 
recesses d, shown also in section to the left. These, by the move- 
ments of the valve, control the exhaust ports e, which connect with 
the main exhaust on the side of the chest. The throw of the valve 
is caused by the admission of a little live air through the small 
grooves j, j, to the ends of the chest, this air being alternately 
discharged by the much larger auxiliary ports /, /. These open 
into the cylinder through the auxiliary ports g, exhausting at 
each stroke into the annular recess of the piston and thence into 
one of the square ports h, which lead to the main exhaust. The 
ports g are bushed with composition metal rings, shaped at the 
lower end to fit closely upon the piston. 

As shown in the cut, the piston is in position to begin its forward 
stroke; the valve has been thrown to the right and is admitting 
air to the rear end of the cylinder. The drills are designed to 
run at the high speed of from 450 to 500 strokes per minute, accord- 
ing to the air pressure. Other details of construction are shown, 



3i6 



COMPRESSED AIR PLANT 



including the cradle or 
shell, section of the rifle- 
bar ratchet and its pawls, 
a new form of chuck and a 
^'dust allay er." 

The chuck comprises a 
heavy wedge and half 
bushing, with a long bear- 
ing surface which grips the 
bit shank firmly by means 
of the U-bolt. A tap with 
a hammer loosens the 
wedge, permitting rapid 
changing of bits. When 
worn, the chuck bushing 
is set up by a liner, to keep 
the bit in the axis of the 
machine. 

A specialty of this drill 
is the ''dust allayer " 
(shown in detail, plan, and 
elevation), which is at- 
tached to the air tap by 
a nipple and cup, forming 
a ball - and - socket joint. 
It is, in effect, an ejector, 
drawing water from any 
convenient source, by 
means of a small quantity 
of compressed air led from 
the throttle. By the same 
air the water is sprayed 
forward into the mouth of 
the drill hole. 

Stephens & Son build 
piston valve drills of the 




COMPRESSED AIR ROCK DRILLS 3I7 

following sizes: if in., 2 in., 2} in., 2^ in., 3 in., 3 J in., and 3 J 
in.; also tappet valve drills of 2f in., 3 J in., and 3 J in. diameter 
of cylinder. 

Holman Drill. There are two forms of this English machine, 
built at Camborne, Cornwall: the air- or spool-valve drill, made 
in 2, 2 J, 2 J, 2|, 3-J, 3 J, and 3I inch sizes; and the tappet drill, of 
^h sly Sh ^nd 3f inch diameter of cylinder. The 2 and 2 J inch 
drills are of light weight, intended chiefly for stoping in thin veins. 

Spool-valve Drill. In a few of their details the smaller sizes 
of this type differ somewhat from the larger, though the general 
design is substantially the same in all. Fig. 142 illustrates the 
sizes from 2{ to 2| inch. The movements of the spool-valve i, 
which control main air and exhaust ports of the usual form, are 
caused as follows: Below each end of the valve-chest, and com- 
municating from chest to cylinder, is a short vertical air port, 
containing a coned or taper seating. In each of these ports is a 
pair of steel balls, 4, 5, the former of which controls the auxiliary 
port 7. Both balls are under the pressure of the spiral spring 8. 
The seat is so shaped that the lower and smaller ball 5, will project 
slightly into the cylinder, whenever permitted to do so by the 
position of the annular recess 6, around the middle part of the 
piston. Hence, by each stroke of the piston, the lower ball 5 
receives a slight upward blow from the inclined shoulder of the 
recess. This lifts the larger ball 4, also; thereby opening the 
auxiliary port 7, and placing the corresponding end of the valve- 
chest in communication with the main exhaust 3. Owing to 
the pressure of the air occupying the opposite end of the chest, 
the spool-valve is then reversed, to prepare for the next stroke of 
the piston. Ball valves are well adapted for this service. They 
are not liable to breakage, and, as they receive a slight rotary 
motion from each blow of the piston, the wear tends to be equalized, 
thus keeping them round and preventing leakage between the balls 
and their seats. 

Tappet-valve Drill (Fig. 143). This differs from most of the 
American-made drills of the same type in the shape of the tappet 
and piston. The oscillations of the tappet 10, are caused by the 



COMPRESSED AIR ROCK DRILLS 319 

enlarged middle portion, ii, of the piston; the upper arm of the 
tappet engaging with the D-valve 12. A spring, 13, holds the 
valve on its seat. 

German Drills. A number of good machine drills are made 
in Germany, most of them designed on the spool- or air-valve 
principle, and differing chiefly in the smaller details from the 
American and English machines already described. Amongst 
them are those of: P. Hoffmann, Eiserfeld; R. Meyer, Miihlheim- 
Ruhr; Froelich and Klupfel, Unter-Barmen; Duisburger Maschi- 
nenbaii, Duisburg; H. Flottmann and Co., Bochum; Freimann 
and Wolf, Zwickau and the Kiizel drill, of R. W. Dinnendahl, 
Steele. Several of these machines, for example, the Duisburger, 
use hollow bits and a water jet. 

Some repetition would be necessary to describe these drills 
and details are therefore omitted. A brief description is given 
below, however, of the ''Triumph" drill, built by H. Schwarz and 
Co. (Ruhrthaler Maschinen-f abrik) , Miihlheim-Ruhr. In being 
valveless and in the mode of distributing and exhausting the air, 
this machine is interesting in having a strong resemblance to the 
recent "hammer" drills, and also to the old Darlington drill, for 
many years well-known in Great Britain. 

"Triiimph" Drill (Fig. 144). The cut shows a longitudinal 
section, together with a transverse . section through the rotating 
ratchet and pawls. Air is admitted' by the two-way throttle valve 
c, on top of the cylinder, entering thence the annular port d. At 
the beginning of the stroke, as showri in the cut, d is in communica- 
tion with an annular recess e, near the forw^ard end of the piston; 
whence the compressed air passes through /, which is one of four 
longitudinal ports in the body of the piston, to the rear end of the 
cylinder. The forward stroke then takes place, the air in front 
of the piston being exhausted through the ports j. As the piston 
advances, the exhaust ports j are covered by the solid part 
of the piston, thus cushioning the end of the stroke. When the 
annular recess e, in the piston, comes opposite the ports j, the air 
exhausts from the back end of the cylinder. At the same time, 
the annular recess h, near the rear end of the piston, comes into 



;20 



COMPRESSED AIR PLANT 




S 

3 



COMPRESSED AIR ROCK DRILLS 32 1 

connection with the air inlet d, thus admitting live air through the 
four longitudinal piston ports, one of which, i, is shown in section. 
These ports conduct the air to the forward end of the cylinder and 
the stroke is reversed. 

The parts are so proportioned that the air acts at full pressure 
throughout only about one-third of the forward stroke and then 
expands. Should the piston strike the cylinder head, the shock 
is absorbed by the spring a. In the front head is a stuffing box, 
the gland of which is held in place by the cap h. 

Contrary to the usual construction of rotating devices, the 
rifle-bar k is solidly screwed into the rear end of the piston and 
engages with a correspondingly rifled nut m, in the ratchet g. The 
outer end of the rifle-bar reciprocates in the closed tube f, which 
is connected with the cylinder by the small passages x and y. As 
a result, live air acts on the entire area of the rear end of the piston, 
including the area of the rifle-bar. The tube f, and connecting 
passages, serve incidentally for the better distribution of oil on both 
sides of the ratchet and rifle-nut. 

Temple-Ingersoll " Electric-Air " Drill. A discussion of electric 
rock drills — employing the term in the usual sense — would here 
be inappropriate. The "Electric- Air" drill is unique in the mode 
of combining both systems of power transmission and in no w^ay 
belongs to the class of electric-driven drills, which for years have 
been brought out from time to time, but which as yet have not 
given wholly satisfactory results.* 

The Temple-Ingersoll machine (Fig. 145) comprises three 
parts: a drill, and an air pulsator which is driven by an electric 
motor. Both pulsator and motor are mounted on a small, flat- 
wheel truck, close to the drill and connected with it by two short 
lengths of hose. As shown in the cut, the drill differs in many 
respects from the ordinary rock drills. The cylinder is of larger 
diameter and the short piston, with packing rings, somewhat 
resembles the piston of a steam engine. The drill is carried in a 
supporting and guiding shefl, mounted on a column or tripod, and 

* It may be pointed out that some resemblance to this machine is traceable in 
the design of the "Pneumelectric Coal Puncher" (Chapter XXII). 




w 



a 



COMPRESSED AIR ROCK DRILLS 323 

is provided with a feed screw and rotation device. It is valveless 
and has no buffers, springs or side rods. The pulsator is in effect 
a small, vertical, duplex, single-acting compressor, with cranks 
set at 180°; the crank-shaft being driven through single re- 
duction gearing from the armature shaft of the motor. From 
the pulsator cylinders one length of hose passes to the back end 
of the drill cylinder, the other to the forward end. These connec- 
tions serve as ports for admission and return of the compressed air. 
There is no exhaust. The air circuit is closed, the same air being 
used over and over again. Thus the speed of the stroke of the 
drill depends on the speed of the motor, and this is varied by a 
controller, operated through a cord by the drill runner. If a 
direct current motor be used, it is designed for three speeds; or 
an alternating current motor, single speed, may be employed if 
desired. 

The pulsator runs at a low air pressure, only a small degree 
of compression being necessary for transmitting the power and 
acting as a spring between pulsator and drill. Incidentally, the 
air cushions the drill piston at the ends of the stroke. Leakage of 
air from joints and past the pulsator pistons is provided for by a 
compensating valve (not shown in the cut), which is adjusted to 
maintain a practically constant pressure in the air circuit. When 
the pressure falls below the limit, the valve opens automatically 
and admits a little more air. This air is compressed by the 
differential area between the two parts of the piston in the first 
cylinder, until the normal working ^'"essure is restored. Lubrica- 
tion of the pulsator cranks and pistons is effected by the ''splash" 
method, the lower part of the crank-case being partly filled with 
oil. A portion of the oil is atomized and carried with the air into 
the drill cylinder. 

The "Electric- Air" drill is made in four sizes as shown in 
Table XXIX. 

In working capacity these machines correspond approximately 
to the 2 in., 2| in., 3 in., and 3^ in. sizes of the Sergeant air drill, 
described previously. The voltage recommended is 220. For 
alternating current the standard motors (which are stronger and 



324 



COMPRESSED AIR PLANT 



Table XXIX 



No. 


Diameter 
Cylinder, 
Inches. 


Stroke, 
Inches. 


Strokes 

per 
Minute. 


Weight of 
Drill Un- 
mounted, 
Pounds. 


Weight of Pulsator 
AND Motor, Pounds. 


Approxi- 
mate 
H.P. at 




D. C. 


A. C. 


Pulsator 

for One 

Drill. 


3-C 
4-D 

4-E 
5-C 


3f 
4i 
4f 
Si- 


61 

7 
7 
8 


475 
415 
440 
400 


119 
223 
228 
299 


645 

883 
1050 


370 
545 
928 
820 


3 



simpler than those for direct current) are three-phase, 25, 30, 50, 
and 60 cycle. Direct current motors, wound for 440 or 500 volts, 
may be used, but these pressures are unnecessary and are dangerous 
for underground service. 

Air Pressure for Machine Drills. The evidence adduced from 
recorded tests shows conclusively that a low air pressure is un- 
economical. Both the force of the blow and the number of 
strokes per minute fall off, resulting in a marked decrease in the 
footage of hole drilled. While it is probable that drilling in soft 
rock does not require so high an air pressure as for hard, it is 
found on the whole that the best results are obtained by a pressure 
of from 70 to 80 lbs. Practice of late has tended toward the 
use of higher pressures, up to 90 lbs. or even more; but, grant- 
ing that more work in some kinds of rock may be done by em- 
ploying a heavier pressure than, say, 80 lbs., the life of the drill is 
shortened and the cost of repairs increased. The customary 
nearly uncushioned blow, under a heavy air pressure on hard 
rock, becomes very destructive to the machine, and the bits them- 
selves do not stand so well. They are dulled sooner and are 
more apt to chip. 

The influence of air pressure, as well as the questions relating 
to air consumption per drill, are further illustrated by a number 
of tests made several years ago in the South-African gold district.* 

* J. B. Carper and others, Mechan. Engineers Assoc, of the Witwatersrandy 
1904. (Abstract in Mines and Minerals, Sept., 1904, p. 64.) 



COMPRESSED AIR ROCK-DRILLS 



325 



The rock in which the tests were made was red granite, a large 
block of which was embedded in concrete. A quarry bar was 
used for mounting the drills. All holes were drilled vertically, 
with abundance of water. Two receivers were employed, with 
a combined capacity of 757 cu. ft., the pressure for each run 
being raised by the compressor to 80 lbs., after which the 
receiver was shut off. A single machine at a time was operated, 
the run continuing until the receiver pressure dropped to 70 lbs. 
The drill was then stopped, and the depth and diameter of hole 
measured. Similar runs were successively made with pressures 
from 70 to 60, 60 to 50 lbs., etc. The capacity of the receiver, 
in terms of cubic feet of free air, was calculated for each in- 
dividual run and pressure, correction applied for temperature, 
and the air consumed based on the volume of free air at 70° F. 
and 24.8 ins. of barometer (equivalent to an altitude of 5,000 ft.). 
Eliminating several results of such runs as indicated erratic 
behavior of the drills, probably due to being in poor condition, 
a test of 13 drills, 3 J ins. diameter with 3-in. bits gave the 
following averages : 

Table XXX 





Air Pressure, Pounds. 




80-70 


70-60 


60-50 


50-40 


40-35 


Linear inches drilled per min 


1-3 
124. 

95-3 
13-3 


I.I 
117. 
106.4 
14-8 


I.O 

100. 
100. 
13.8 


0.6 
70. 
116. 4 

15.0 


60. 


Cu. ft. free air per minute 


Cu. ft. free air per linear in. of hole 

Ditto per cu. in. of hole 


120. 
16.6 







Each run occupied about 6 minutes. Some of the average 
results are not consistent, and the individual figures of course 
showed still greater variations. These were due to a variety of 
causes, such as lack of uniformity of the rock, differences in 
temper and sharpness of bits and, in a measure, the personal 
equation of the drill-runners, each of whom " was selected by the 
agent of the maker of the drill." The rather lengthy paper from 
which these data are taken includes many tables, giving details 



326 COMPRESSED AIR PLANT 

of the tests of machines of different makers, and is to be recom- 
mended for the thoroughness with which the work was summa- 
rized. Among other points, the importance of the question of air 
pressure is clearly demonstrated. 

Consumption of Air. By reason of the irregularity of the 
work of machine drilling, and the fact that in mining or other 
rock-excavation work a number of drills are always operated by 
the same compressor plant, few figures are available as to the 
actual air consumption of a single machine. Average figures, 
however, are the only really useful ones. It is customary to base 
the duty on the consumption of free air per minute, the quantity 
necessarily depending on the size of the machine, air pressure 
supplied by the compressor, character of the rock, and the pro- 
portion of the total time actually occupied in drilling. It is 
evident that the compressor capacity for a single machine is 
greater than the average required for a number of machines. 
With a large number, the delays to which each is subject, for 
setting up or shifting, changing bits, stoppages caused by the bit 
sticking in the hole, etc., make it improbable that all of them 
will be in simultaneous operation, save in rare instances; hence, 
the average allowance of air for each may be reduced. Mo- 
mentary or occasional peaks in the load on the compressor, when 
an unusual number of drills happen to be working simultaneously, 
may be disregarded; or at least need not be provided for by in- 
creasing the compressor capacity. 

Rock-drills of different makers, even when of the same diam- 
eter of cylinder, vary in their consumption of air and reliable 
figures are not easily obtained. Table XXXI, showing the volume 
of free air per minute reqitired for one drill, is based on a com- 
parison of the statements of several manufacturers, checked by a 
few recorded tests. It may be taken to represent, within reason- 
able limits of error, the results of actual practice for machines in 
good order. No allowance is made for the preventable loss of air 
in leaky pipes, nor for frictional loss of pressure in transmission 
(see Chapter XVI). 



COMPRESSED AIR ROCK-DRILLS 



327 



Table XXXI 

Cubic Feet of Free Air per Minute Consumed by One 
Drill at Sea-Level 



Gauge 






Diameters of 


Drill 


Cylinder in Inches. 






Press- 
















ure. 




























2 


2K 


2% 


2% 


3 


3/8 


3x'b 


3K 


zv^ 


3^8 


4?4 


5 


60 


58 


63 


70 


82 


90 


97 


100 


105 


114 


118 


135 


155 


70 


62 


72 


80 


92 


104 


112 


115 


118 


130 


135 


152 


174 


80 


70 


80 


88 


103 


115 


125 


130 


135 


142 


153 


173 


205 


90 


7« 


«7 


95 


115 


128 


137 


141 


148 


i(>5 


173 


194 


222 


100 


«5 


96 


108 


126 


140 


151 


155 


161 


176 


184 


210 


250 



When a number of drills are operated by the same plant, the 
compressor capacity for furnishing the total average quantity of 
free air required per minute, at sea -level, may be found approxi- 
mately by the following table of multipliers: 

Table XXXII* 



Number of drills. . . 


I 


2 


3 


4 


5 


6 


7 


8 


9 


10 


Multiplier 


I 


1.8 


2-7 


3-4 


4.1 


4-8 


5-45 


6.x 


6-7 


7-3 






Number of drills. . . 


II 


12 


15 


20 


25 


30 


35 


40 


50 


60 


Multiplier 


7-8 


8.4 


10.3 


12.8 


iS-i 


17.3 


19.7 


22.0 


26.5 


30-5 



The required capacity of the compressor is found by multiply- 
ing the cubic feet of free air per minute consumed by a single drill 
(as given in Table XXXI), by the multiplier corresponding to the 
number of drills operated (Table XXXII) . 

It will be understood from what precedes that the figures 
in the tables cannot be taken as exactly applicable to all cases. 
Several other modifying factors may here be summarized : 

(i) The Kind of Work. The time required to set up the drill 
depends greatly on the shape of the working, whether a tunnel or 
drift, a shaft, stope, or open cut. If the floor and roof, or the side 
walls, of a mine opening are irregular or loose, much time may be 

* Based on comparison of several tables given by manufacturers. 



328 COMPRESSED AIR PLANT 

lost in shifting the machine and setting it up, according as it is 
mounted on column or tripod. 

(2) Character of the Rock. This also influences the consump- 
tion of air. In hard rock the rate of advance in drilling is slower 
than in soft, so that the machine makes longer continuous runs. 
Less total time is occupied in shifting and setting up for drilling 
the successive holes of a round, and the consumption of air per unit 
of time is therefore greater. Though this increase is partly offset 
by the fact that the bits are more quickly dulled in hard rock and 
must be changed at shorter intervals; still, in very hard ground 
the machines may be kept running with but few and short in- 
termissions. In soft rock, on the other hand, though the actual 
speed of drilling is greater, there are apt to be more frequent 
delays due to rifling of the hole and sticking or " fitchuring " of 
the bit. On the whole, for hard rock it is advisable to provide a 
greater compressor capacity than is given in the tables. The 
compressor will then be able to run at a slower speed, thus avoid- 
ing excessive heating in cylinder and receiver. In general, the 
time actually occupied in drilling will vary for each machine 
from, say, 4 to 6 hours out of an 8-hour shift. 

(3) Physical Condition of the Drill. The importance of this 
matter may be overlooked. The figures given are for new 
machines, or those in thoroughly good order. More air is con- 
sumed by old drills, whose valves and pistons are so worn that 
they do not fit closely. Even in the case of drills in fair average 
condition, this is clearly shown by the fact that the exhaust, in- 
stead of being short and sharp, is nearly continuous. A large 
allowance must be made for old machines. 

If definite values could be assigned to these different items, 
estimates of air consumed per drill could be made in conformity 
with any given set of conditions. To do this is manifestly im- 
possible, but a few general data relative to averages for an entire 
shift's work have been put on record by Messrs. J. E. Bell and 
L. L. Summers, as the result of a series of experiments (Mining 
and Metallurgy, Feb. ist, 1901). For a 3-in. drill, the volume of 
free air required per shift of 8 hours is as follows, the gauge 
pressure being 100 lbs. ; 



COMPRESSED AIR ROCK-DRILLS 329 

Table XXXIII 



Elevation. 


Cubic Feet of Free Air. 




Per Shift of 8 Hours. 


Per Minute. 


Sea-level 


25,000 to 42,000 
30,000 " 49,000 
35,000 " 60,000 


52.1 to 87.5 
62.5 " 102.0 
73.0 " 125.0 


5,000 ft 

10,000 ft 



These figures include all deductions, for whatever cause, cov- 
ering delays and stoppages as well as the actual drilling time. 

Taking the various allowances into account, and applying 
them to Tables XXXI and XXXII, the following results, obtained 
in an elaborate test made at the Rose Deep Mine, Johannesburg, 
South Africa,* will be found in fairly close agreement with 
what precedes. The average number of drills (Ingersoll-Ser- 
geant), of several different sizes, kept in operation during the 6- 
hour test, was calculated to be equivalent to 30.9 drills, 3 J in. 
diameter of cylinder. The average duty per drill was 4 ft. 
5l^ ins. of hole per hour (diameter of hole not stated). 
Average air pressure, 69.83 lbs. Free air used per drill per min- 
ute, 81.08 cu. ft. It is fair to assume that most of these drills were 
more or less worn, or at least not in perfect condition. Accord- 
ing to the tables, the average free-air consumption for 30.9 drills 
should have been about 68 cu. ft. per minute, or about 15 per cent, 
less than that shown by the test. This difference is accounted for 
in part by the altitude above sea-level. It may be added that the 
horse-power per drill developed in the steam cylinders of the com- 
pressor was 12.72. But as the work done during the 6-hour 
test was approximately equal to that usually accomplished in 
8 hours of regular work, the actual horse-power per drill under 
normal conditions in this mine may be taken as 12.72 X-| = 9. 54. 
The air piping in this case was known to be remarkably free 
from leaks. 

Another test run, on 75 drills, 3 J in. diameter, was made 

* L. I. Seymour, South African Association of Engineers, 1898. 



^^O COMPRESSED AIR PLANT 

about 5 years ago at the Champion Iron Mine, Mich.* At 
78 lbs. normal gauge pressure the average air consumption for 
the day shift, throughout a period of i month, was 67.1 cu. ft. of 
free air per minute. The air piessure usually dropped consider- 
ably, however, when work was in active progress. According 
to the tables, 75 drills should have used an average of about 58.5 
cu. ft. of free air per minute, or 13 per cent, less than shown by 
the test. 

Efficiency of Machine Drills. Though it is a well-known fact 
that compressed-air drills are uneconomical machines in con- 
sumption of power, it is difficult to reach definite conclusions as to 
their efficiency. The actual useful work — employing this term in 
its ordinary mechanical sense — done by a machine drill in making 
a hole of given depth and diameter in a rock of given hardness, 
toughness, and general physical character cannot be determined 
absolutely. All that is really known is that the drill requires a 
certain volume of air per minute, which has been furnished by the 
expenditure of a certain average indicated horse-power at the com- 
pressor. Comparative figures of work done, in terms of speed of 
drilling in a given rock and per cubic foot of free air consumed, are 
often published and are useful as far as they go. In fact, this is 
the only practical basis for estimating their efficiency. But, even 
in this sense, the propriety of accepting the results obtained, as 
accurately representing the value and efficiency of machine drills 
as compared with various forms of air or steam engines, may 
well be questioned. 

In their operation rock-drills differ greatly from other com- 
pressed-air machines, because the personal element of the skill 
and experience of the drill-runner exerts so important an influence 
upon the amount of work accomplished, and because the rate of 
drilling is so greatly modified by the physical and mineralogical 
character of the rock, together with the purely adventitious occur- 
rence of cracks, slips, and fissures. A skilful drill-runner will in- 
evitably do more work per shift, under the same conditions, than 
an inexperienced man, and he will make a faster rate of advance 

* Engineermg and Mining Journal, May i8th, 1905, p. 937. 



COMPRESSED AIR ROCK-DRILLS 33 1 

in a rock with which he is specially familiar than if called on 
to operate a machine in rock that is new to him. 

Therefore, though mechanical efficiency, pure and simple, is 
the basis upon which machines in general are compared, in the 
case of compressed-air drills it is not the only consideration, 
nor is it the most important. Their efficiency of operation is 
subordinate to the attributes of strength, simplicity of construc- 
tion, portability, durability, ease and readiness with which re- 
pairs may be made and capacity for work in terms of depth of 
hole drilled per unit of time. They must be capable of with- 
standing hard and often unintelligent usage. The strong point 
of compressed-air drills is their ready applicability in the special 
and peculiar field of work for which they are designed. In 
possessing a cylinder, piston, and valve, the drill roughly re- 
sembles a steam engine, but there the likeness ceases. Severe 
shock and vibration are essential accompaniments of its work. 
No fly-wheel is admissible, or other means of storing up and 
equalizing the power, and the service demanded from the rock- 
drill is therefore totally different from that performed by ordinary 
engines. 

The low theoretical efficiency of the compressed-air drill is 
due mainly to the fact that air is admitted to the cylinder prac- 
tically throughout the full stroke. As a consequence, the valve 
motion bears a strong resemblance to that of many of the single- 
cylinder, direct-acting pumps. Expansive use of the air to any 
extent is neither advisable nor practicable, both because of the 
undesirability of introducing complexity of mechanism in ma- 
chines subjected necessarily to rough usage and because of the 
difficulty of adapting cut-off gear to the variable length of stroke 
required. Owing to the nature of its work, the drill cannot be 
kept always at full stroke. While in operation it is often neces- 
sary to feed the machine so far forward that the actual length of 
stroke may be little more than one inch, and the valve motion 
must still be capable of reversing promptly. A sharp, quick 
reversal of the stroke is essential. The useful work is done on 
the forward stroke, in striking the blow. If the valve be thrown 



332 COMPRESSED AIR PLANT 

too soon, the stroke of the piston will be shortened; if too late, 
the piston may strike the cylinder head. For these reasons, it is 
impracticable with machine drills to attain the economy resulting 
in other air motors from using the air expansively. Incidentally, 
the use of air at full stroke is of some advantage, because, in ex- 
hausting at high pressure, the exhaust air issues from the port at 
a high velocity, and its force, combined with the development of 
some heat from friction, in a measure prevents troublesome ac- 
cumulation of ice, in case the air is moist. The freezing, if any, 
is at least confined to the exterior portion of the exhaust port, 
whence it is easily removed. 

In dry, dusty mines it is generally found that the tappet valve 
gives the better service. When a compressed-air drill is not in use, 
and disconnected from the air hose, dust and grit are likely to enter 
through the ports, passing thence into the valve chest and cylinder 
on resuming drilling. The wear and consequent looseness in the 
fit of the moving parts thus caused is apt to have a more un- 
favorable effect on the operation of the spool than the tappet 
valve. Leakage of air past the valve or piston prevents proper 
action of the auxiliary ports, not only producing irregularity in 
reversal and shortening of the stroke, but diminishing the drill's 
efficiency. It is true that the tappet valve involves the use of one 
extra part and, in case of the three-arm tappet, breakage is not in- 
frequent. But while the spool valve is strong and reliable, ex- 
perience indicates that in dusty mines at least the mainte- 
nance cost of the spool-valve drill is higher than that of the 
tappet drill. 

The maximum force of blow is attained by drills which take 
air throughout the full forward stroke, i.e., without cut-off, and 
the best drills are designed to work m this way. On the forward 
stroke the valve is not reversed until the blow is delivered, the 
exhaust being free, with but little back-pressure on the piston. 
Cushioning was formerly made a feature of some rock-drills, with 
the idea of reducing shock, but it is now recognized that the 
efficiency is increased by delivering an uncushioned blow. It is 
possible for a drill so designed to strike too heavy a blow in very 



COMPRESSED AIR ROCK-DRILLS ^^^ 

hard rock, but the remedy then is to feed the drill-head down, so 
as to work with a shorter stroke. 

On the back stroke cushioning is desirable, to ease the re- 
versal and prevent injury by the piston striking the rear cylinder 
head. The back-stroke cushion is produced by cutting off the 
exhaust before the end of the stroke. Only enough power needs 
to be developed on this stroke to overcome the resistance due to 
the weight of the moving parts, and the frequent tendency for the 
bit to stick fast in the hole. 

In ordinary machine drills, the piston speed should not be too 
great — say, not much over 350 to 375 strokes per minute. The 
relative speeds of stroke do not constitute a proper basis for the 
comparison of efhciencies. To give effect to the blow, the weight 
of the moving parts must be relatively great, and a very high 
speed would be attended by excessive wear and breakage. These 
conclusions do not apply, however, to the numerous small air 
hammer drills which have come into favor in the past few years. 
(See Chap. XXL) The hammer drill strikes a light blow, some 
of them at the rate of 2,000 to 3,000 or more strokes per minute. 
Thus the weight of the moving parts is small, and the inertia 
moderate. 

Drill Repairs. There are a number of good machine drills 
on the market, whose relative merits it is difficult or impossible 
to determine. In choosing a drill the question of repairs is of 
great importance. But very little useful information concerning 
this point is available; and in fact such data could only be obtained 
by operating under the same conditions, and for a considerable 
period of time, drills of several different kinds. Such opportunities 
rarely exist. It is not usual, nor generally advisable, to use different 
makes or more than two sizes of drill in the same mine or surface 
excavation, since this practice involves the necessity of keeping 
dupUcate spare parts for each. 

The item of repairs depends largely upon the experience and 
character of the drill runner. A careful man will treat his machine 
with intelligent consideration. He will set it up properly, to 
reduce the risk of getting out of alignment as the hole is deepened; 



334 COMPRESSED AIR PLANT 

and, if the bit should stick ("fitcher"), will keep his temper and 
refrain from striking unnecessarily heavy blows on the drill head 
or chuck. A fitchered bit may often be loosened by slacking the 
clamp bolts, thus allowing the machine slightly to shift its position. 
The serious abuse to which machine drills are frequently subjected 
may be reduced by efficient supervision on the part of foremen and 
shift-bosses.* 

It is desirable, in every piece of machinery, that there shall be 
as few moving parts as possible. But, in a machine performing 
the severe work of a rock-drill, and often operated by incompetent 
men, there is another consideration. Even when run with care, 
wear is rapid and breakages are frequent. The maker of a machine 
drill, therefore, has the problem before him of producing a drill 
consisting of as few parts as practicable; designing it, at the same 
time, so that those parts which experience shows to be specially 
liable to wear, derangement or breakage, may be replaced readily, 
cheaply and without the necessity of discarding perhaps a much 
larger piece, with which the broken part may be connected. A 
properly equipped repair shop is a matter of importance for length- 
ening the life of machine drills. New cylinders may be bored out 
to fit worn pistons, and new pistons fitted to old cylinders. Drill 
runners should not be encouraged to tinker their machines under- 
ground. If repairs or adjustment be necessary, the drill should be 
sent at once to the shop. 

Records of Work. If it were possible to secure approximately 
complete records, tabulations would best show the comparative 
speeds of drilling in different rocks and ores. But the local con- 
ditions are obviously so variable that n© summary comparison 
can justly be made. In the following pages I have attempted 
to elucidate the subject by citing a number of cases. A few of 
these are the results of observations by the author; most of them 
have been obtained from engineers in the field and from the de- 
tailed notes taken by mining students, who have studied at different 
mines under the direction of the author. Unless otherwise stated, 

* For a good article on the operation of machine drills see Mining and Scientific 
Press, 1905, November 4, p. 308; November 11, p. 329. 



COMPRESSED AIR ROCK-DRILLS 335 

the "total time" in each case includes all ordinary delays in drilling, 
for changing bits, etc., except the time occupied in setting up the 
machine. This usually takes from 15 to 30 minutes. 

San Pedro Copper Mine, N. M. A 5 ft. x 7 ft. cross-cut, in 
very hard quartzite and limestone; lo-hour shift; one Ingersoll- 
Sergeant 3-inch drill; average air pressure, 60 lbs. 

Heading advanced in one month 38.3 ft. 

Number of drill-hours 357 

Total number of holes drilled. 248 

Total number of feet of hole 770 

Average depth of hole 37 in. 

Average number of feet of hole drilled per hour 2.16 

A 5 ft. X 7 ft. drift in the same mine, in quartzite and vein 
matter, with the same drill and air pressure, was advanced 46.6 ft. 
in one month : 

Number of drill-hours 207 

Total number of holes drilled 113 

Total number of feet of hole 430 

Average depth of hole 45 in. 

Average feet of hole per hour 2.08 

In the above records all delays for breakage, setting up the 
machine, and changing bits, are included. 

Table XXXIV. 



Hole. 


Depth, 


Total Time, 

Including 

Delays. 

Minutes. 


Net Drilling 
Time. 

Minutes. 


Feet per Minute. 


Feet. 


Total Time. Ne 


t Time. 


I 


4-7 


50 


38 


.094 


124 


2 


5 


7 


63 


■ 56 


.090 


102 


3 


6 


4 


70 


56 


.091 


114 


4 


4 


9 


48 


37 


.102 


132 


5 


5 


3 


44 


39 


.120 


136 


6 


6 





57 


50 


.105 


120 


7 


6 





90 


72 


.066 


083 


8 


5 





56 


44 


.089 


113 


9 


5 





45 


37 


.III 


135 


10 


4 


5 


47 


■43 


.096 


105 


II 


6 


3 


30 


27 


.210 


231 


12 


5-0 


43 


31 


.116 


161 


Averages 


5-4 


53-6 


44 


.107 


129 


Average inch 


es per 


minute 






I .29 ] 


■55 



336 COMPRESSED AIR PLANT 

Snowstorm Mine, Idaho. A drift in hard quartzite; 8-hr. shift; 
Ingersoll-Rand 3j-inch drill; air pressure, 80 lbs. (Table XXXIV). 

Two Tunnels for U. S. Reclamation Service, North Yakima, 
Wash. Tunnel section, 7 ft. x 7 ft. 3 in., in fissured, hard but 
blocky basalt; one Wood 3-inch drill; 8-hour shift. 

Average advance per month 143 . 5 ft. 

Number of feet of hole per shift 29 

Average depth of hole drilled per hour 3 • 62 ft. 

Average cost of drill repairs per shift 58 cents. 

Michigan Copper Mine, Rockland, Mich. Stoping in fairly- 
hard amygdaloidal and brecciated vein rock; Rand 3j-inch drill; 
60-65 lbs. air pressure; 8-hour shift. 

Depth of holes 6 to 8 ft. 

Total depth of hole drilled in i month (26 days) , 554 ft. 

Average depth drilled per shift 37 ft. 

Average depth drilled per hour 4 . 63 ft. 

At the same mine the following observations were made on 
the drilling of individual holes, in drifting and stoping: 

Table XXXV. 



Hole. 


Depth. 
Inches. 


Total Time, 

Including 

Delays. 

Minutes. 


Net DriUing 

Time. 

Minutes. 


Average Inches per 
Minute. 




Total Time. 


Net Time. 


(Drifting.) 

J 
2 

3 
4 


32 
19 

53 
52 


25 

14-5 
37 
48.5 


22.5 
6 

14 
22 


1.28 
1.30 
1.40 
1.07 


1.4 
3-1 
3-7 

2-5 


Averages 


36 


31.2 


16. 1 


1.26 


2.7 


(Stoping.) 

I 
2 
3 
4 
5 
6 

9 


74 

70 

76 . 

66 

76 

81 
64 
68 


62 
36.5 

51-5 
35 

53 


52-5 
26 

61 -75 

29 

36.5 

33-5 

23 

19 

20 


I 
I 

I 


2 
9 

4 
3 

3 


1-4 
2.7 
1 .2 

2-3 

2 .1 

2-3 

3-4 
3-5 
1-7 


Averages 


72 


47.6 


33-5 


1.6 


2-3 



COMPRESSED AIR ROCK-DRILLS 



337 



A Lead Mine, near Flat River, Mo. Stoping and drifting in 
rather hard, tough, bedded limestone, interstratified with bands of 
shale; formation has numerous cleavage planes, generally parallel. 
Sullivan drills are used; U. S. 2j-inch for stoping and U. C. 21- 
inch for drifting; air pressure at drills, 75 to 85 lbs. Average 
depth of hole, 7 feet. Average footage drilled per 8-hour shift, 
39 feet. 

The Waugh drifting, stoping, and sinking drills, Nos. 8 and 
8 D, with hollow steel, were tried at this mine, but were found to 
be unsuited to the conditions. They made, respectively, 36.1 
ft., 43.7 ft., and 22.3 ft. per 8-hour shift. 

Federal Lead Co., Flat River, Mo. Breast -stoping in fairly 
hard and tough limestone, with 2j-inch SulHvan drills. Air 
pressure at drills, 75 lbs. Average depth of holes, 6 ft. 8J in. 
The drills were new; work done by contractors. In direction, 
nearly all of the 25 holes ranged from about 45° above, to 30° 
below, the horizontal (three were nearly vertical). 



Table XXXVI. 



Number 
of Run. 


Number 
of Holes. 


Total Depth 
Drilled. 


Total 
Time. 


Net Drill- 
ing Time. 


Per cent, 
of Shift 
Actually 
DriUing. 


Drilled per 
Minute, 

Net Time, 
Inches. 


Feet. 


Inches. 


Hrs. 


Mins. 


Hrs. 


Mins, 


I 
2 

3 

4 


4 
8 
6 

7 


19 

57 
42 
48 


ID 

10 
8 


3 

7 
6 

7 


18 
24 

47 


I 
4 
3 
3 


06 

48 


48.3 
513 
47-9 

47-5 


2.05 
2.77 
2.23 
2.56 


Av. depth of hole, 


6 


^ 


Average speed of all holes, 


2.46 



Mount Hope Iron Mine, Wharton, N. J. Underhand stoping 
in solid, magnetic ore; Sullivan 3-inch drill; air pressure, 75 to 
85 lbs. 

ist hole, 94 ins. deep, in 68 mins. = 1.38 in. per minute. 
2d hole, 66 ins. deep, in 53 mins. = 1.24 in. per minute. 
Time includes changing bits and other incidental delays, but 
not setting up the machine. 

Detroit Copper CoJs Mines, M or end, Ariz. Stoping with 



33^ 



COMPRESSED AIR PLANT 



Ingersoll-Rand '' Sergeant C-24," 2|-inch drill, in ore of varying 
hardness and mineralogical character, as stated below; average 
air pressure, 75 lbs. 

Table XXXVII. 



Number 


Aggre- 
gate 
Depth, 
Feet. 


Average 
Depth, 
Feet. 


Total 
Time, 
Hours. 


Average Speed of 
Drilling. 


Character of Ore. 


of Holes. 


Per Hour, 

Feet. 


PerMim:te, 
Inches. 


80 
91 

74 


367 
417 
238 


4-59 
4-58 
3.21 


54 

59-5 

88 


6.80 
7.01 
2.70 


1.36 
I .40 
0.54 


Soft porphyry. 
Rather soft quartzite. 
Very hard, tough, 
quartzite. 



The above figures represent six weeks' work. 

El Oro Mining and Railway Co. [ 1 ;- , • a • t h '11 

The Mexico Mines of El Oro. \ 
ing speeds, in stoping in very hard, tough quartz, and for develop- 
ment work (drifting, cross-cutting, and raising) , in quartz, andesite, 
and hard, black shale or slate. Drills used: Ingersoll-Rand 
'' Sergeant C-24," 2f-inch and ''Sergeant A-86," 2|-inch; air 
pressure, from 70 to 75 lbs. 

2|-inch drill: average speed of drilling, including changing 
bits but not shifting and setting up, 1.5 in. per minute; average 
footage per 8-hour shift, 35 ft. 

2j-inch drill: average speed (as above), 1.2 in. per minute; 
average footage per shift, 30 ft. 

Wahana Iron Mines, Nova Scotia. Stoping in red hematite 
(hard-ore type); Sullivan 3-inch, spool-valve drill; 70-75 lbs. air 
pressure; lo-hour shift. 

Depth of holes 6 to 8 ft. 



One Month 
372 



One Month 

Number of drill shifts 340 

Average feet of hole drilled per shift 58.6 69 

Average feet of hole drilled per hour 5.8 6.9 

Average cost of drill repairs per shift 45 .4 cents. 

Pennsylvania Copper Mine, Butte, Mont. Stoping; ore chiefly 
a compact granitic and quartzose gangue, with streaks and bunches 
of chalcocite, moderately hard; Rand 2|-inch drill; 8-hour shift. 



COMPRESSED AIR ROCK-DRILLS 

Table XXXVIII. 



339 



Hole. 


Depth, 
Inches. 


Total Time. 
Minutes. 


Net Drilling 

Time, 

Minutes. 


Average Inches per 
Minute. 




Total Time. 


Net Time. 


I 

2 

3 

4 


62 
66 
66 
64 


30-5 
46.5 

59 


22 
27 

38-5 
32 


2 .00 
1.42 
I .12 


2.80 
2.44 
1.66 
2 . 10 


Averages 


64-5 


45-3 


30 


I-5I 


2.25 



Highland Boy Mine, Bingham, Utah. In a 10 ft. by 10 ft. 
drift, fairly hard, compact limestone, with Rand 3j-inch drill, the 
average drilling speed per hour, for several observations, was 2.43 
feet of hole, including all delays except setting up. One round of 
13 to 14 holes, 4 ft. deep, are drilled and blasted in this drift in 
one 8-hour shift. For stoping in pyritic ore, with a 34-inch Rand 
drill, a round of four 4-ft. holes are drilled in 3 J hours, or at the 
rate of 5.1 ft. per hour. 

Vekol Gold Mine, Pinal Co., Ariz. Results of a drilling contest 
in a 4J by 6 ft. drift, in solid blue limestone; Sullivan 2f-inch 
spool-valve drills; 105 lbs. air pressure; 8- hour shifts. 

Table XXXIX. 



Shift. 


Time for 

Setting Up, 

Minutes. 


Number 

of 
Holes. 


Feet 
Drilled. 


Average 
Depth of 
Hole, Feet. 


Total Work- 
ing Time. 


Average Speed 

of Drilling, 

Feet per 

Hour. 




Hours. 


Mins. 


I 
2 

3 
4 

5 


26 
16 

19 
21 

17 


II 
II 
12 
10 
12 


94 

95 

102 

85 
92 


8.5 
8.6 

8.5 

1-1 


7 
6 

7 
7 
7 


54 
25 
15 
10 
02 


II. 9 

15.2 

14 
II. 8 

13-1 


Totals 


99 


56 


468 




35 


46 




Averages 


19.8 


II .2 


93-6 


8.36 


7 


09 


13.2 



The drift was advanced at the average rate of 6 ft. 2 in. per 
shift. This extremely rapid work is interesting in showing what 



340 



COMPRESSED AIR PLANT 



can be done under the stimulus of the spirit of rivalry produced by 
a "drilling contest." 

Wolverine Copper Mine, Michigan. The following results of 
drilling in the characteristic amygdaloidal vein matter of the 
Keweenaw copper district are abstracted from a paper by W. R. 
Crane {Engineering and Mining Journal, Sept. 8, 1906, pp. 438-9). 
Rand drills were used, of 3 and 3j-inch diameter of cylinder. 

Table XL. 



Kind of Work. 



Drifting \ 
10 holes f 
Drift stoping ) 
9 holes S 

Cutting out stope 
6 holes 

Raise stoping ) 
9 holes ) 



Averages 



Aver- 
age 
Depth, 
Feet. 



5-76 
5.60 

7-50 
6.50 



Total Aver- 
age Drilling, 
Time per 
Hole. 



Min. 



52 
41 
48 

52 



23 
47 

40 
12 



Net Average 

Drilling, 

Time per 

Hole. 



Min. 



36 
24 

29 
41 



Sec. 



44 

04 

54 

GO 



Average Inches per 
Minute. 



Total Time. 



1-31 
I .60 

1-85 
1.50 



1-56 



Net Time. 



3.01 
I .90 



2 .40 



Table XLI. — (Drifting). 



Hole 


Depth, 
Feet. 


Total Time, 
Minutes. 


Net Time, 
Minutes. 


Average Speed, 
Inches per Minute. 




Total Time. 


Net Time. 


I 

2 

3 

4 

5* 

6 


5-5 
5-5 

5-5 

li 

6.0 


31 
32 

148 
50 


26 

27 
30 
31 
100 
40 


2.13 
2 .06 
1.78 

1-73 
0.50 
1.44 


2.54 
2.44 
2 .20 
2.13 
0.72 
1.80 


Averages 


5.66 


56 


42 


1. 61 


1.97 



* Hole No. 5 was seriously delayed by " fitchering " of the bits. Omitting it, the 
average speed of the other five holes is 1.83 in. per minute for total time and 2.22 in. 
per minute for net time. 



compressed air rock-drills 
Table XLII.— (Storing). 



341 











Average Speed, 


Hole. 


Depth, 


Total Time, 


Net Time, 


Inches per Minute. 




Feet. 


Minutes. 


Minutes. 


Total Time. 


Net Time. 


I 


7-5 


46 


38 


2 .00 


2-37 


2 


7-5 


54 


45 


1.66 


2 .00 


3 


6-75 


64 


57 


1.26 


1.42 


4* 


7 


115 


87 


o-n 


0.96 


5 


8 


45 


38 


2.13 


2-53 


Averages 


7-35 


65 


53 


1.56 


1.86 



* Hole fitchered. Omitting it, the average speeds are respectively 1.76 and 2.08 
in. per minute. 

Portland Gold Mine. Cripple Creek, Colo. Sloping in very 
hard phonolite-breccia, with IngersoU 2i-inch drill; conditions 
difficult. 

Table XLIII. 











Average Speed, 


Hole 


Depth, 


Total Time, 


Net Time, 


Inches per Minute. 




Inches. 


Minutes. 


Minutes. 


Total Time. 


Net Time. 


I 


18 


26 


22 


0.69 


0.82 


2 


38 


28 


24.25 


I 36 


I 


57 


3 


25 


46 


42 


0-54 





60 


4 


24 


2>2> 


31 


0-73 





80 


5 


30 


45 




0.66 






6 


?>?> 


23 




•44 






7 


18 


26 


• • 


.69 






8 


38 


36 




1.05 












(4 holes) 




(4 holes) 


Averages 


28 


Z2, 


29.8 


0.89 


0-95 



The above record is below the average for stoping in this mine. 
Where the ground is more favorable, eight 3-ft. holes are usually a 
fair day's work. 

In a drift of the Cresson Mine, also at Cripple Creek, observa- 
tions of two holes, 53 inches deep, drilled by a 2j-inch IngersoU 



342 



COMPRESSED AIR PLANT 



drill, showed the high speeds of: 2.2 and 3.0 in. per min. for 
total time; 3.1 and 3.65 in. per min. for net drilling time. 

Conclusions. The average speed of drilling shown by the 
sixteen examples given in the preceding pages is 6.2 feet per hour. 
In general it may be said that the duty of a standard 3-inch machine 
drill, in rock or ore of average hardness, ranges from 40 to 50 feet 
per 8-hour shift, including all ordinary delays for setting up and 
changing bits. For very hard, tough ground, the speed is often 
much lower; while considerable more than 50 feet per shift may 
be made when the conditions are favorable, and also in drilling 
deep holes, for which fewer set-ups are required. The cost per 
foot of hole is obviously extremely variable, ranging from say 8 
cents in easy ground, and where wages are low, up to 25 cents, 
when the conditions are adverse. 

For moderately soft ground, not requiring holes of large 
diameter to contain the necessary quantity of powder, the smaller 
sizes of machine drill — from 2 in. to 2 J in. — are usually preferable. 
Their first cost, as well as air consumption, is less than for large 
drills, and they may usually be operated by one man. These 
small machines are specially useful for stoping in rather thin veins. 
But, for hard ground, and as a rule in shaft sinking, tunneling, 
cross-cutting and similar work, the 2| in., 3 in., and 3 J in. sizes 
are best. For deep holes in large surface excavations, still heavier 
drills are often necessary — up to 3J in., or even larger. 



CHAPTER XXI 

COMPRESSED AIR HAMMER DRILLS 

The principles of the hammer drill were first applied in pneu- 
matic riveting hammers, and tools for chipping, rough chiseling 
and miscellaneous machine shop work. Their earliest employ- 
ment in mines was for cutting hitches for timbers, block-holing, 
and other small work, where no great depth of hole is required.* 
In recent years they have rapidly grown in favor, as the result of 
the introduction of improved designs, and are now competing with 
the larger and heavier reciprocating machines for all kinds of 
comparatively shallow drilling, like quarry work and stoping in 
thin veins; also for sinking shafts and winzes where it is desirable 
to use small and light machines. 

All the different makes of hammer drill are built in several 
sizes and patterns. The smaller machines weigh only i8 lbs., and 
are held in the hands of the operator, a D-shaped handle being 
provided for convenience (Fig. 148). These light drills are de- 
signed for block-holing and sinking, and in general for holes 
directed below the horizontal. Many of the larger sizes are 
designed with an air-feed standard, for overhead work (Fig. 151), 
and may or may not be mounted on a light column. These, too, 
are easily carried from point to point and set up and run by one 
man. Still others, like the Kimber, of South Africa, are, in 
diameter of cylinder and weight, about equal to the small sizes of 
reciprocating drills. Lastly, the Leyner hammer drill, which 
may be said to form a class by itself, is built of the same sizes as 
the average standard machines of the reciprocating pattern. 

* The writer believes that the Franke hammer drill, brought out in Ge^nany 
in 1891 or 1892, was the first hammer drill used in mining. (See Zeitschr. 
jiir das Ber g-Hiitten-und Salinenwesen, Vol. 41, p. no.) It weighs 16 lbs. and 
strikes several thousand blows per minute. 

343 



344 COMPRESSED AIR PLANT 

General Construction. In the hammer drill, the bit does not 
reciprocate. The shank of the bit projects into the forward end 
of the cylinder and is struck a rapid succession of blows by the 
piston, which acts as a hammer. When in operation the cutting 
edges of the bit are in constant contact with the bottom of the hole, 
except during the slight rebound caused by each blow of the 
hammer. Many of these machines are valveless, the functions 
of the valve being performed by the reciprocations of the hammer 
or piston. Others, like the Leyner, Sullivan, Climax, and one of 
the types of the Inger soil-Rand, are provided with spool-valves. 
With the exception of the Leyner hammer drill and one or two 
others, no attempt is made to introduce automatic rotation of the 
bit. Rotation in all the smaller machines is effected by hand, 
the operator turning the whole machine back and forth on its 
axis, by means of the handle. The bit shank is made octagonal, 
generally fitting loosely in the chuck socket, which is of the same 
shape. To keep the hole round and reduce the chances of rifling, 
the bit is commonly of the star shape, with six (sometimes eight) 
radial cutting edges. This construction brings the cutting edges 
so close together that even if several successive blows are made 
in the same position of the bit, the ridges of rock between the edges 
are broken away. 

As the bit does not reciprocate, it is evident that, for holes 
directed downward and more than a few inches deep, some auto- 
matic means must be provided for removing the drill dust or 
sludge and keeping the bottom of the hole clean, otherwise much 
of the useful effect of the blows of the hammer would be lost. To 
accomplish this, a hollow bit is generally used, a small hole being 
bored longitudinally through its center. By injecting a jet of water, 
the drillings are displaced and the bit is kept cool. The same 
result is attained by a jet of compressed air, which produces a low 
temperature on expanding. As the speed of stroke of hammer 
drills is great, the cooling of the bit in dry holes is important. 
The air jet is obtained by exhausting through the bit, at each 
stroke, part or all of the air from the cylinder. When using 
water, it may be led to the chuck through a special passage in the 



COMPRESSED AIR HAMMER DRILLS 345 

cylinder; or a small tube is inserted longitudinally through the 
center of the piston, to the inner end of the bit. The water is 
supplied by gravity, or under artificial pressure (as for the Leyner 
drill). 

The small hand hammer drills are fed simply by keeping the 
bit pressed firmly against the bottom of the hole. In these, much 
of the shock and vibration are absorbed by a handle of rubber 
hose. Some of the larger machines of all the principal makes are 
provided with an automatic air-feed device. This adapts them 
for general service, and specially for drilling holes at a steep up- 
ward angle, as required in overhand stoping and in making 
"raises." Usually the automatic feed consists of a light telescopic 
standard, screwed into the back of the cylinder. It is supplied 
with compressed air, which keeps the drill fed up to its work as 
the hole is deepened. Incidentally this device furnishes an air 
cushion, relieving the operator from much of the annoyance 
caused by vibration. These machines may be mounted on a light 
column, when desired, for breast-stoping, drifting, and similar 
work. Details of the air-feed are given hereafter. 

Leyner Drill. Although this well-known machine must be 
classed with the "hammer" drills, its construction and operation 
are widely different from the numerous small, light machines 
working on the hammer principle. It occupies the same field as 
the standard drills of the reciprocating type; is designed for 
mounting on tripod or column and is provided with automatic 
rotation of the bit. 

The most important drill of this make is commonly designated 
as the "Water" Leyner. Fig. 146 shows the longitudinal section. 
The cylinder, 17, is .carried in guides in the shell, i, and is fed by 
the feed screw and nut, 23 and 21, and feed crank, 26; 13 is the 
hammer, 7 the chuck, in which the drill shank is held by the 
chuck-key, 6; the chuck parts, including the buffer, 3, being re- 
tained in place by the cap, 2. The air valve, 15, is of the spool 
or piston type, controlling the main ports, and being itself actuated 
by a system of small auxiliary ports, opened and closed by the 
reciprocations of the hammer. 



COMPRESSED AIR HAMMER DRILL 



347 



Rotation of the bit is effected as follows: The rifle-bar, i8, 
which is controlled by the pawl, with spring and plunger, 31, 32, 
and ^Tf, engages with the rifle-nut, 16, screwed into the hollow 
rear-end of the piston or hammer. This causes rotation of the 
hammer on each back stroke. The forward, smaller end of the 
hammer is fluted, and engages with an internally fluted bronze 
nut, 9, in the rear end of the chuck. Thus the chuck, holding 
the bit, is caused to rotate with the hammer. These parts are 
shown separately in Fig. 147. The drill bit, also shown in this 




Fig. 147. — Rotation Device of "Water" Leyner Drill. 



cut, is of special form, being not only hollow, for the passage of 
the water into the drill hole, but provided also with two lugs, 
by which it is locked in the chuck. 

The water supply, already referred to, is furnished under 
pressure from an i8-gallon steel tank, accompanying the drill 
and connected to it, at 37, by a length of hose. Another hose 
conveys compressed air from the main to the tank. Water is 
thus-forced from the tank through the water tube, 39, which passes 
through the rifle-bar and hammer, in the axis of the machine, and 
is delivered into the hollow bit, when the latter is locked in the 
chuck ready for work. A back buffer, for easing the shock, if 
the hammer should overrun its stroke, is shown at 20, with its 
plate, 19; II and 12 are the front cylinder buffer and buffer ring. 
At 41 and 42 are leather washers for making air-tight the joints 
between chuck and hammer. 

The use of the water jet undoubtedly increases the rate of 
drilling; it keeps the hole clean, so that the bit is not cushioned 
— as it may be with other machines — by an accumulation of stiff 
sludge in the bottom of the hole. In solid rock of uniform texture 



348 COMPRESSED AIR PLANT 

the Leyner does good work, but is not so successful in fissured, 
seamy formations, where a machine drill bit is likely to stick or 
"fitcher." Moreover, to obtain the best results from this machine, 
more careful and intelligent handling is required than for most of 
the reciprocating drills. 

The "Water" Leyner drill is made in two sizes. No. 2 has 
a 2j-inch cylinder and 9-lb. hammer, the drill head, bare, weighing 
130 lbs., or, with all fittings, boxed for shipment, 205 lbs. No. 3 
drill, bare, having a 3-inch cylinder and i3i-lb. hammer, weighs 
155 lbs.; including fittings, 230 lbs. Without the leg-weights, 
the tripod weighs about 200 lbs.; 6J-ft. column, 275 lbs., and 
water tank, 70 lbs. 

Another Leyner drill is the "Rock Terrier," a small machine 
for stoping and similar work, handled by one man. It is made in 
two patterns: the "dry," of which the drill head, bare, weighs 52 
lbs., and the "water" pattern, bare, 54 lbs. The hammer weighs 
2| lbs.; 6-foot column, with clamps, 85 lbs., and tripod with shell, 
but without leg-weights, 46 lbs. 

Still another pattern of the Leyner drill is the No. 5, a light- 
weight but strongly built machine, specially designed for stoping. 
It differs essentially from the larger machines of the same make: 
(i) in having no rifle-bar, with its accessories, for the automatic 
rotation of the bit; (2) in being provided with an automatic air- 
feed cylinder or standard, such as is found in a number of the 
machines described in the following pages. The valve-chest, 
containing a spool-valve somewhat similar to that of the "Water" 
Leyner, is attached to the side of the main cylinder. 

Hardsocg Wonder Drill. Fig. 148 is a longitudinal section of 
one of the smaller sizes, provided with a D-handle; Fig. 149 shows 
a drill with the air-feed attachment. 

The operation of the hand drill will be understood by referring 
to the first cut, in which the hammer is shown at the end of its 
forward stroke. Compressed air is admitted at the nipple, 2, to 
which is attached the air valve and hose connection to the main. 
From 2 the air enters the annular recess, 3, and acts constantly on 
the shoulder, 13, of the hammer. At the beginning of the forward 



COMPRESSED AIR HAMMER DRILLS 



349 



stroke, the 2-way ports, 5 and 6, passing through the head of the 
hammer, are opposite the recess, 3. Air is thus admitted into and 
behind the hollow hammer, and the area presented to the air 
pressure being much greater than the area of the shoulder, 13, 
the forward stroke is made. Just before the hammer strikes the 
bit,-the ports 5 and 6 reach the annular space 8, in the front end 
of the cylinder. The air is exhausted into 8, whence part of it is 
discharged directly into the atmosphere by the exhaust port, 4, 
the remainder passing through the hollow bit into the bottom of 
the drill hole. (When drilling in soft rock, and especially when the 
presence of moisture produces a pasty sludge, all of the exhaust 




Fig. 148. — Hardsocg Wonder Drill, with D-Handle. 



air may advantageously be discharged through the bit, by plugging 
the port, 4.) The exhaust having taken place, the back stroke is 
made by the constant pressure of the inlet air on the annular 
shoulder, 13, of the hammer. This drives back the hammer until 
the ports 5 and 6 are again brought opposite the recess 3, thereby 
admitting air behind the hammer. 

The drill is fed forward by hand, and is rotated by the handle, 
12, the bit shank being octagonal. To take up vibration, the 
handle is made of rubber tubing, held in place by studs, as shown. 
x\n air-feed cylinder (see second cut), may be screwed into the 
rear end of the drill cylinder, by removing the large plug and the 
rubber handle. The rotation is then effected by manipulating 
the handles, 11. When using the air-feed, the admission port, 2, 
is plugged, the air entering from the feed cylinder through the 



350 



COMPRESSED AIR PLANT 



longitudinal port, 9. These drills strike from 1,800 to 2,000 
blows per minute, at 90 to 100 lbs. air pressure. 

Fig. 149 is a section of one of the larger machines, provided 
with the air-feed and designed for drilling up-holes. The feed 
cylinder, which is from 3 to 4 ft. long, is "broken" in the cut, to 
reduce the length. When at work, the drill rests on the pointed 
end of the feed cylinder, as nearly as possible under the place 
where the hole is to be drilled, and the operator steadies and holds 
the machine in the exact position required. Holes at a considerable 
angle to the vertical may thus be readily drilled. 

Air is admitted at 15. A small portion of it passes through 
the port 29 to the rear of the piston, 17, by which the machine is 
fed forward automatically. Air for the drill cylinder passes from 




Fig. 149. — Hardsocg Air-Feed Stoping Drill. 



15, round the feed piston rod, to the central port, 18; and thence, 
through ports 19 and 20, to the annular recess, 21, surrounding 
the hammer. When the transverse port, 22, comes opposite to 
the recess, 21, air is admitted to the hollow hammer, causing the 
forward stroke. At the end of the stroke the exhaust takes place 
by the ports 22 and 23, opening to the atmosphere at 24. As in 
the hand drill, the back stroke is produced by the constant pressure 
of live air on the annular shoulder of the hammer. 

The impact of the hammer is received by the anvil or tappet 
block, 25. The front head, 27, in which is set the chuck, 26, 
is attached to the cylinder by the sleeve, 28. This sleeve has 
differential threads, for locking it firmly in position. 

The hand drills, like that in Fig. 148, are made in four sizes, 
weighing from about 19 to 45 lbs. Air-feed machines (Fig. 149), 



COMPRESSED AIR HAMMER DRILLS 



351 



in a number of sizes and patterns, weigh unmounted from 35 to 
65 lbs.; or, including their light columns and arms, from 95 to 
160 lbs. In Cjuarry work, for making straight rows of holes, as 
in getting out dimension stones, air hammer drills may be mounted 
on a horizontal bar, supported at each end by a pair of adjustable 
pointed legs. This mounting is similar to, though much lighter 



I7v 




tb^iKMt==ffJi 



E^S^---^^&I 




Fig. 150. — Murphy Hammer Drill, with D-Handle, for Sinking. 



than, that used in quarry service for ordinary reciprocating rock 
drills. 

Murphy Drill. This resembles the Hardsocg in general con- 
struction, though with differences in details. 

Fig. 150 illustrates the design of the cylinder and ports of the 
latest model of the "sinker" drill, provided with D-handle. The 
hammer, 13, is bored out from the rear end, and has a transverse 
port, 16, near its forward end. Compressed air enters at the hose 
connection 17 and, on raising the throttle valve, 33, passes to the 
annular recess, 18, in the cylinder 20. The throttle is set in a 
branch pipe, 32, screwed into the cylinder, and has a gland, 35, 
lock nut, 37, packing and washer, 36. (Incidentally this valve 



352 



COMPRESSED AIR PLANT 



casing serves as a second handle, assisting the operator in support- 
ing and rotating the machine.) On entering the recess, iS, the 
air acts on the shoulder, 19, thus driving back the hammer. At 
the end of the back stroke, the port, 16, is in connection with the 
recess, 18, and compressed air is admitted to the hollow hammer. 
The forward stroke is thus made, as in the Hardsocg drill. When 
the port, 16, reaches the larger diameter of the cylinder, at the end 
of the forward stroke, the air at the rear of the hammer is exhausted 
through the port, 14. The stroke is then reversed by the constant 
pressure on the shoulder, 19. The back-head, 23, of the cylinder 
is screwed on like a cap. It has a lock-key 26, and a washer, 25, 
which serves also as a buffer. A simple bushing, 10, holds the 
bit, which has a collar as shown, but neither chuck-clamp nor 
bolts are used. This bushing is keyed in place by the bolt, 22, 
in the ring, 21, encircling the forward end of the cylinder. 

Live air is admitted to the hollow drill steel, through the hose 
connection, 40, from the throttle valve to the auxiliary valve, 47. 
This is set, with gland and washer, in the valve casing, 42, attached 
to the back cylinder head by the nipple, 43. From the auxiliary 
valve the air passes through a small diagonal port to the rear end 
of the tube, 38, which is held in the back head by the gland, 39, 
and communicates with the hollow steel through the forward end 
of the hammer. By manipulating the auxiliary valve, the operator 
regulates as required the quantity of air passing to the drill hole. 

The Murphy sinker drill is made in two sizes, Nos. 3 and 5, 
weighing 55 and 65 lbs., and using respectively about 50 and 60 
cu. ft. of free air per minute. 

The air-feed stoping drill, for overhead work, is built in two 
models, two sizes of each; weighing, without column and arm, 
from 85 to 95 lbs. Model L is shown in Fig. 151. Air enters 
by the hose connection, 17; passing thence through the passages, 
18 and 19, to the drill cylinder. Admission is controlled by the 
hand valves, 7, 7. From 18 a small quantity of air is admitted 
through port, 16, to the upper end of the feed cylinder, causing the 
inner tube, 50, carrying the pointed end, 61, to slide outward. 
Part of the feed air passes through the central port, 35, to the 



354 COMPRESSED AIR PLANT 

annular space between the inner and outer tubes, and acts on the 
shoulders shown at each end. The standard is thus lengthened, 
feeding the drill forward automatically as the hole is deepened, 
and keeping the bit firmly pressed against the bottom of the 
hole. The stuffing box for the telescoping tubes is shown at 64, 65 
and 66. 

In this machine the branch pipe, 32, instead of containing the 
throttle, is used merely as a handle for rotating the drill. The 
hammer does not strike the bit directly, but delivers its blows on 
the tappet block, 12. In other respects the cylinder and its 
appurtenances are substantially the same as those of the Murphy 
sinker drill, described above. Owing to the mode of construction 
of this model, the air hose connection turns with the machine, so 
that in rotating it the operator cannot make complete revolutions, 
but must turn it back and forth. Care must be taken to set up 
the drill so that no dirt will enter the lower joint of the stuffing- 
box. This joint, however, is protected by its position from the 
cuttings falling from the mouth of the drill hole. 

The other model, H, of the stoping drill has the hose connection 
attached to the outer tube of the feed cylinder, and the stuffing-box 
is at the end nearest the drill cylinder. In this position the 
stuffing-box is more exposed to wear from the drill cuttings, but 
as it is not near the floor the machine can be set up anywhere, 
without a supporting block or plank. Moreover, in rotating the 
drill the operator can make complete revolutions, if desired. 

Sullivan Hammer Drills. Unlike the machines already de- 
scribed these have a valve for distributing the air. Figs. 152 and 
153 show longitudinal sections of the latest pattern of hand drill, 
with hollow bit, and illustrate the valve motion and arrangement 
of ports, which are essentially the same for all the models. In 
the earlier forms a solid spool-valve was used; in the present 
design the valve is cup-shaped and works on the spindle, Z, which 
is a part of the valve-box cap, as shown by the cuts. The section 
in Fig. 153 is taken at about 45° from that in Fig. 152, for the 
purpose of exhibiting all the longitudinal ports. 

Air is admitted, through the hose connection to the port Y 



356 COMPRESSED AIR PLANT 

and recess, B, simply by pressing the handle of the drill. This 
forces down the throttle plug, U, to the position shown. When 
the handle is released, the plug rises automatically, under constant 
pressure from the small live-air lead, E, and closes the port Y. 
The hammer, H, is shown on its forward stroke. Air enters from 
B, through several longitudinal passages, C, to the annular port, 
D; thence, by the rear annular valve groove, to port F, and finally, 
by a series of other longitudinal passages, G, through the ports, S, 
to the cylinder. 

While the drill is in operation, the port, J, communicating with 
B, is occupied by air at constant pressure. Previous to the de- 
livery of the blow of the hammer, its groove, I, comes opposite to 
the port, J, thereby putting the port, K, in communication w^ith 
J, and admitting live air through P to the forward or annular end 
of the valve, V. As the area of this end exceeds that of the other 
(which is subject to constant air pressure from the inlet B), the 
valve is reversed. This puts the annular valve port or groove, D, 
in communication with the long reverse port, X W. Meanwhile, 
the hammer completes its stroke on momentum, while air is being 
admitted at W to the forward end of the cylinder, preparatory to 
reversing. The air in the rear end of the cylinder, which caused 
the forward stroke, exhausts through the series of ports, S, leading 
to the corresponding longitudinal passages, G, thence to the 
annular valve port, F, and finally reaches the atmosphere 
through M. 

The back stroke is caused by the pressure of the air on the 
annular shoulder, T, of the hammer. When the small, forward 
end of the hammer passes out of the contracted portion of the 
cylinder, the air which was used to hold the valve upward passes 
through port N to the exhaust opening, O. Simultaneously, the 
air which caused the back stroke also exhausts through O. The 
live air in B now acts on the upper or smaller end of the valve, 
driving it forward and reversing the hammer. 

This type of Sullivan drill is made in tw^o si^es. The ''plug" 
drill, for light work, has a i-5/16-inch cylinder and weighs 18 lbs. 
Either hollow or solid steel may be used; if solid, part of the air 



358 



COMPRESSED AIR PLANT 



is exhausted through a hose connection to an annular rider on the 
bit shank, by which jets of air are directed into the hole. Another 
size, with if- inch cylinder, weighs 30 lbs., uses hollow steel, and 
bores holes up to if -inch diameter, with an air consumption (at 
100 lbs. pressure), of about 25 cu. ft. free air per minute.,, 

The air-feed drill is shown in Fig. 154. The valve' motion is 
the same as described above, except in the arrangement|fpf^ the air 
admission passages and the f)osition and manipulation of the 
throtde, A. The latter is set transversely in an eccent|ric sleeve, 
G. As the sleeve is turned by the handle, around the drill, it 
carries with it the valve. The eccentric inner surface of G causes 
the valve to take the different positions necessary for controlling 
the admission ports. First, it opens the port to the feed cylinder, 
E, and then, by turning the handle farther, opens the port leading 
to the constant pressure chamber back of the main valve, as in 
the hand drill. The throttle plug is held against the eccentric 
surface of the sleeve by the small lead of live air, as shown. 

The feed cylinder, E, is similar to those of the other makes of 
air-feed drills, already described. Live air from the admission 
port acts on the piston, C, which is attached to the inner telescoping 
tube, D. When in operation, the whole drill is rotated back and 
forth by the handle, F. The diameter of cylinder is 2 inches; 
weight of machine, 70 lbs. ; maximum diameter of hole drilled, 2 
inches; free air consumption (at 100 lbs. pressure), 35 to 40 cu. 
ft. per minute. 

IngersoU-Rand " Imperial " Hammer Drills. This company 
makes hammer drills of two distinct classes, namely: the "Im- 
perial," which is valveless, and the "Crown" drill, provided w^ith 
a spool-valve. Each of these classes comprises machines of several 
different sizes and weights, both of the D-handle and telescope- 
feed types. 

The "Imperial" hammer drill, with D-handle, types M V-i 
and M V-2, is shown in longitudinal section by Fig. 155. Its 
general construction is similar to that of the valveless machines 
already described. Air is admitted at the connection, 13, into 
which is screwed the nipple, 46, shown in the figure in the lower 



COMPRESSED AIR HAMMER DRILLS 



359 




360 COMPRESSED AIR PLANT 

part of the cut. To 13 is attached a short hose connection, with 
stop-cock, 59, and nipple, 48, for joining to the hose leading from 
the air main. The hollow piston 5 has a shoulder near the rear 
end, against which the air pressure is constantly acting, and a 
series of 6 slot-shaped ports near the forward end. In the cut the 
piston is shown as having completed its stroke, in which position 
the air is being exhausted through the piston ports and thence 
to the atmosphere. The return stroke is caused by the constant 
pressure on the shoulder of the piston. 

Rotation of the drill is provided for by a straight handle, 
screwed or clamped by a gland to the cylinder. In this design 
the clamp bolt and nut are shown at 8 and 9. The shank of the 
bit is held in a bushing, 69, which is made in two forms, to receive 
either hexagonal or cruciform steel. Solid steel is used for these 
machines, which are designed specially for shallow "plug holes,'* 
both for mining and quarry service. They may be employed also 
for miscellaneous drilling, in connection with many engineering 
and contracting operations. 

Fig. 156 illustrates the " Imperial" air- or telescope-feed hammer 
drill, type MC-12. The hammer, 5, is shaped differently from 
that of the M V style, described above, but its action is similar. 
An anvil-block, 22, transmits the blow of the hammer to the bit 
shank. The bit bushing, 67, is held in the cylinder head by the 
cottar-bolt, 86. The construction of the telescope-feed is clearly 
indicated in the cut and will be understood by reference to the 
descriptions already given of air-feed hammer drills. Compressed 
air is admitted to the machine at the connection, 36, controlled 
by an air cock. From the upper end of the telescope feed tube 
the air passes through several radial ports in the rear cylinder 
head, and thence into the cylinder through longitudinal ports, as 
shown. For rotating the machine a straight handle, 48, is screwed 
into the cylinder casting. 

The ''Imperial" drills weigh from 42 to 65 lbs., unmounted; 
approximate air consumption, at pressures from 60 to 100 lbs., 
13 to 57 cu. ft. per minute. Travel of telescope-feed, 20 inches; 
total length of drill, with feed run in, 48J and 50 inches. Number 




Oh 



362 



COMPRESSED AIR PLANT 




a. 



P^ 



COMPRESSED AIR HAMMER DRILLS 363 

of blows per minute, at 60 lbs. pressure, from 990 to 1,320; at 
100 lbs., from 1,120 to 1,560. 

IngersoU-Rand " Crown " Hammer Drills. These machines 
are provided with spool-valves and are known as types HA, H B, 
and H C, according to size and weight. General longitudinal 
sections of the telescope air-feed and D-handle styles are shown 
in Figs. 157 and 158. 

The valve motion is illustrated by the diagrams. Figs. 159 and 
1 60. At the beginning of the stroke the valve, F, and hammer, 
H, are in the positions indicated in Fig. 160. Air enters port A 
from the hose connection. The inner end of A being now closed 
by the hammer, the air passes through ports B and C to the valve 
chest. Part of the air entering through C escapes by the small 
port, D, thus reducing the pressure in the forward end of the valve 
chest, E, below the working pressure. The pressure in E holds 
the valve in the position shown, so that the air entering at B passes 
around the valve into the port G, and thence into the rear end of 
the cylinder. This drives the hammer forward. While the 
hammer is making its stroke, part of the air in the front end of the 
cylinder exhausts through port I, and part by the longitudinal 
port, J, around the valve, F, and finally through port K to the 
atmosphere. 

Referring to Fig. 159: as the hammer approaches the end 
of its forward stroke, the annular groove^ L, is opposite port M, 
thus allowing live air from port A to pass into the longitudinal port, 
O, and thence to the rear end, P, of the valve chest. The pressure 
of this live air being greater than the reduced pressure m chamber 
E (Fig. 160), already referred to, the valve is thrown forward to 
the position shown in Fig. 159. Then air from port B passes 
around the valve and through port J, to the front end of the 
cylinder, thus causing the back stroke of the hammer. During 
this stroke the air in the rear end, of the cylinder exhausts through 
port G, and thence around the valve to the exhaust port Q. While 
on its back stroke the hammer opens port I, allowing all compressed 
air in port O, and in the front end of the cylinder, to escape. This 
exhausts the air from the chamber, P, of the valve chest, and the 



3^4 



COMPRESSED AIR PLANT 




COMPRESSED AIR HAMMER DRILL 



365 




366 COMPRESSED AIR PLANT 

constant reduced pressure at the other end of the chest (E, Fig. 160) 
causes the valve to reverse and assume its original position, thus 
completing the cycle of operation. The H A-14 type of this drill 
is provided with a short hose, from the valve chest to a "dust- 
hose nozzle" encircling the bit at the mouth of the drill hole. 

The telescope-feed machines weigh from 40 to 80 lbs., un- 
mounted, and consume, according to the makers, from 18 to 60 
cu. ft. of free air per minute, at working pressure from 50 to 100 
lbs. The travel of the telescope-feed is from 18 to 24 inches; 
total length of drill, with feed run in, from 50 to 59 inches. Number 
of blows per minute, at 60 lbs. pressure, from 1,050 to 1,200; at 
100 lbs., from 1,300 to 1,400. 

Waugh Hammer Drill. Fig. 161 shows in plan and longitudinal 
section type 8-C, of this drill, provided with air-feed standard for 
use in stoping. 

Admission of air is controlled by the taper throttle valve and 
handle 12, the valve being in its open position. From the hose 
connection 13, the air follows the path indicated by the arrows 
drawn in full lines, through the hollow throttle to the head of the 
feed cylinder; thence back through a transverse port in the 
throttle, and through longitudinal passages in the valve chest 15, 
to the annular groove A in the chest. The valve 16 has a differ- 
ential shoulder, against which the live air in groove A acts, thus 
throwing the valve back into the position shown in the cut. In 
this position the air passes the forward edge of the valve into the 
main cylinder, and drives the piston or hammer forward against 
the tappet block, 8. During the stroke, after the hammer passes 
port G, the air in front of the hammer is exhausted through port F 
and passages E and D to the annular groove C in the valve chest; 
and thence past the valve through ports H H and passage K to 
the atmosphere, as shown by the broken line arrows. At the 
same time, as soon as the hammer uncovers the port P, live air 
flows from the cylinder through P (and the small dotted port 
forming its continuation) , to the back or larger end of the valve, 
16, thus throwing the valve forward. This shuts off admission 
of air and opens the cylinder space behind the hammer to the 



368 COMPRESSED AIR PLANT 

exhaust, preparatory for the return stroke. The exhaust air 
passes through two round dotted ports, H H, nearly opposite the 
rear end of the valve, to the longitudinal passage K and thence to 
the atmosphere. The exhaust is sufficiently throttled down in 
passing through H H, to keep the valve 16 in its forward position 
by the pressure exerted on its rear or large end. 

In this position of the valve, live air passes as indicated by the 
full Hne arrows, from the annular groove A, through the ports B 
in the valve seat, to the groove C in the valve chest, and thence by 
passages D and E and the port F, to the forward end of the hammer, 
thus causing the return stroke. When, on the back stroke, the 
hammer uncovers the port G, the air exhausts from the front end 
of the cylinder through G and thence to the atmosphere by the 
passage K, as indicated by the broken line arrows. On com- 
pletion of the back stroke, the exhaust pressure acting on the 
larger end of the valve 16 has fallen; so that the live-air pressure 
in the groove A, acting on the small end, throws the valve back 
to the original position (as in the cut). Towards the end of the 
stroke the small end of the hammer enters the hollow valve, and 
there confines and compresses a small quantity of air, thus accel- 
erating the back throw of the valve. 

It may be added that the exhaust air goes through the longitu- 
dinal passage K, and so does not disturb the dust falling from the 
drill hole. 

When the throttle is closed, the inside of the feed cylinder 4 is 
open to the atmosphere through passages Q and R. When the 
throttle is open, the feed cylinder is closed to the atmosphere, and 
at the same time live air is admitted to the feed cylinder through 
a small port in the throttle (not shown), and the drill is carried 
forward to its work. At the same time air enters from the hose 
connection through the hollow throttle and acts on the piston of 
the feed cylinder. 

The stoping drill is made in two sizes, weighing respectively 
50 and 60 lbs., and using solid bits. Drifting drills (8-D and 3-D), 
weighing 75 and 60 lbs., are designed for a column mounting and 
may be used also for sinking. For these two machines, hexagon, 




Clamp Swivel and 
Cradle Sliding Piece 















1 


i^d 


mmi^^^^^^^^^ 




'^yyyy^y^y^^^ 




mmmm \ 




z^^^^^^ 


^m 




^^ \ 






^^^^ 


^^« 


^^S^ 




IP^iii 


k Wa 


terway 



SECTION ON LINE A-B 




SECTION ON LINE C-D 




Waterway from Air 
and Water Tap combined 



Airway to Automatic 
Feed Piston from Tap 




I Water ReguJ 



Fig. 162. — Stepli 



^iisLi^ 




\^W" 



[mperial" Hammer Drill. 



flr- 



V 



COMPRESSED AIR HAMMER DRILLS 369 

hollow steel is used, through which live air or water is delivered 
by means of a 3-way valve. This valve is screwed into the side 
of the drill head, and is opened or closed by a sleeve fitting over 
the rotating handle, 18, so as to be readily manipulated by the 
operator. Water is supplied from a small tank, under pressure 
of compressed air, similar to that described under the Leyner drill. 

Small hand machines, with D-handle, for sinking and for drill- 
ing '^plug" holes, are also furnished by the same makers. They 
weigh respectively 48 and 27 lbs. 

Stephens* "Climax Imperial" Hammer Drill; an Enghsh 
stoping drill, made at Carn Brea, Cornwall. It has a i|-inch 
cylinder, weighs 75 lbs., and is intended to be mounted on a light 
column or bar. In several features of its design the Climax drill 
is in marked contrast to the American hammer drills. 

Referring to Fig. 162, it will be seen that the valve motion 
resembles that of the Climax reciprocating drill, described in 
Chap. XX. Air is admitted by the combined air and water tap 
(detailed section), attached to the side of the valve chest; thence 
passing by the annular recess b, in the piston valve a, through c 
and c\ to the main cylinder ports h, h\ As shown by the cut, the 
valve and hammer are in position to begin the forward stroke. 
The recesses d, d\ in the valve, communicate with the main exhaust 
(not shown) . Air is constantly admitted to both ends of the chest 
by a very small groove t, the valve being thrown by exhausting 
through the much larger auxiliary ports e and /. When the drill 
is in operation, the ports / are alternately brought into communica- 
tion with the annular recess in the hammer, thus releasing the air, 
by way of the square ports 5 and s\ to the main exhaust. The 
ports / are lined with hollow, conical plugs, g, of composition metal, 
which are shaped below to the curve of the hammer. They are 
kept in close contact with the hammer, for preventing leakage of 
air, by the pressure of the valve-chest, when bolted in place. When 
the plugs wear too loose, a thin washer is inserted above them. 

The water for the drill hole is best supplied by gravity, under a 
pressure of say 15 lbs. It enters the combined air and water tap, 
or throttle, already noted, through the passage k, to the transverse 



370 



COMPRESSED AIR PLANT 



port I, in the cylindrical anvil block u, which serves as the drill 
holder, or chuck. Thence the water passes to the hollow bit (see 
the elevation and the '' section on line A B"). The drill may also 
be used, under proper conditions, for "dry" holes, the dust being 
allayed by an external spray from the throttle.* 

The machine has an automatic air-feed. A small piston, p, 
with its rod, is rigidly bolted to the lug q, on the cradle or shell r. 
Air from the throttle passes through the passage j to the feed 
cylinder, thus forcing the entire drill head forward on the shell 
and keeping the bit pressed against the bottom of the hole. After 
the machine has been fed forward 14 inches (the working length 
of feed), the air is shut off and the transverse bolt, shown in the 
plan of the cradle, is slacked. The operator then slides the 
cradle forward on its support under the machine, tightens up the 
bolt on the serrated edge of the cradle, and proceeds with the drill- 
ing. Thus, a total feed of twice the length of the cradle — or about 
28 inches — is obtained without putting in a longer bit. 

Rotation of the bit is effected by hand. The bit is held merely 
by friction in the conical socket of the chuck. Gear teeth are cut 
on the periphery of an enlarged part, 0, of the chuck, engaging 
with which is a smaller gear n (see general plan and the "section 
on line C D"), which is keyed on a spindle passing to the rear of 
the machine and rotated by the handle m. 

Excellent drilling records have been made by this machine, 
both in England and South Africa. 

Makers of Hammer Drills. The hammer drills chosen for 
illustrating the construction and operation of this type of machine 
are believed by the author to be among the best of their respective 
classes. Details of others, equally good, might be given if 
space permitted. A number of these machines are on the market; 
in fact, within the past few years, most of the manufacturers of 
reciprocating rock drills have added to their lists one or more 
styles and sizes of the hammer drill. Arranged alphabetically 
below are the names of most of the drills: 

* This "dust allayer" is described in Chapter XX, under the Climax reciprocat- 
ing drill. 



COMPRESSED AIR HAMMER DRILL 371 

Boyer (Chicago Pneumatic Tool Co., Chicago, 111.). 

Cleveland (Cleveland Pneumatic Tool Co., Cleveland, Ohio). 

Climax (R. Stephens & Son, Carn Brea, Cornwall, England). 

Flottmann (H. Flottmann & Co., Cardiff, Wales). 

Franke (Made in Germany). 

Her (Her Rock Drill Man'f'g Co., Denver, Colo.). 

IngersoU "Crown" (Ingersoll-Rand Co., New York). 

Kimber (Formerly made in Johannesburg, S. Africa, now built 

by the Ingersoll-Rand Co., New York). 

Leyner (J. Geo. Leyner Engineering Works Co., Denver, Colo.). 

"Little Jap" (Ingersoll-Rand Co., New York). 

Murphy (C. T. Carnahan Man'f'g Co., Denver, Colo.). 

Rand "Imperial" . .(Ingersoll-Rand Co., New York). 

Schmucker (Great Western Pneumatic Tool Co., Denver, Colo.). 

Shaw EcHpse (Shaw Pneumatic Tool Co., Denver, Colo.). 

SulHvan (Sullivan Machinery Co., Chicago, 111.). 

Waugh (Denver Rock Drill and Machinery Co., Denver, Colo.). 

Whitcomb (Whitcomb Hammer Drill Co., Rochelle, 111.). 

Wonder (Hardsocg Wonder Drill Co., Ottumwa, Iowa). 

Depth of Hole and Speed of Drilling. In connection with the 
work of hammer drills, the important questions of speed of drilling 
and the practical limit of depth of hole are closely related. Records 
of rapid work are often published and it has been stated, in general, 
that holes can be drilled at the rate of 3 to 4 inches per minute in 
rocks of the granite type, or from 7 to 9 inches per minute in lime- 
stone or ordinary sandstone. Such statements, though well 
substantiated, cannot be taken as general averages. They apply 
obviously to work done under favorable conditions. It is certain, 
also, that, with the air jet and to some extent also with the water 
jet, when the holes are "wet," i.e., directed downward, they do 
not clean so well as depth increases. This begins to be noticeable 
at depths of even 2 or 3 feet, and the speed of drilling materially 
diminishes. Furthermore, as the length and consequent weight 
of the bit increases with depth of hole, the inertia also increases 
and the blows of the light hammer used in most of these machines 
become less and less effective. The deepest holes are made and 
the fastest work done when drilling "uppers" (holes directed at 
a steep upward angle) , in dry rock or ore. The dust and cuttings 
will then run out freely by gravity; the hole is in effect self-cleaning 
and the blows of the hammer are more effective than when the 



372 



COMPRESSED AIR PLANT 



bit is clogged with cuttings, or, in wet rock, with a pasty sludge. 
In drilling dry "uppers," a depth of 5 ^ to 6 feet is probably 
near the economic limit for all except the largest sizes of the 
hammer drills. 

Records of Work. The prefatory remarks, made under this 
heading in Chap. XX. respecting the work of reciprocating drills, 
apply also here. But, since hammer drills are almost invariably 
operated without mounting, no allowance of time for setting up 
the machines is necessary; and as there are no chuck-bolts to 
manipulate, changing bits usually takes only ij to 2 minutes for 
each bit. 

Portland Gold Mine, Cripple Creek, Colo. Stoping in hard, 
phonolite breccia, with Leyner air-feed drill: 

Table XLIV. 



Hole. 


Depth, 
Inches. 


Total Time, 
Minutes. 


Net Drilling 
Time, Minutes. 


Inches Drilled per 
Minute. 




Total Time. 


Net Time. 


I 
2 
3 


24 
37 
17 


25 
135 

7 


16.5 
II 
6 


0.96 
2.74 
2-43 


1-45 
336 
2.82 


Averages 


26 


15-2 


II .2 


2.04 


2.54 



In the same stope, working under the disadvantage of incon- 
venient, cramped positions, 8 holes, aggregating 21 ft. 7 in. deep, 
were drilled in 4 hours, including all delays, or at the average rate 
of 1.08 inches per minute. 

At one of the Michigan Copper Mines the No. 5 Leyner, on 
a test run, drilled 159 holes, aggregating 1,141 feet in 39 shifts: 
or at an average rate of 29.3 feet per shift.* 

Central Tunnel, Idaho Springs, Colo. An average of a number 

* Extract from a letter from the J. Geo. Leyner Engineering Works Co., Dec. 15, 
1909. 



COMPRESSED AIR HAMMER DRILL 



373 



of holes in tough, hornblende schist, drilled by a "Water" 
Leyner machine, showed 2.64 inches per minute, including 
all delays. 

Cresson Gold Mine, Cripple Creek, Colo. Stoping in hard, 
dense trachyte, containing numerous small quartz stringers; 
Waugh air-feed drill: 

Table XLV. 



Hole 


Depth, 
Inches. 


Total Time, 
Minutes. 


Net Drilling 
Time, Minutes. 


Inches Drilled per 
Minute. 




Total Tim-e. 


Net Time. 


I 
2 

3 
4 


36 

51 

54 


16 
18 
22 
26 


14 
14 
17 

22 


2.25 
2.77 
2.32 
2.08 


2-57 
3-57 
3.00 

2-45 


Averages 


48 


20.5 


16.8 


..35 


2 .90 



In this mine the average footage of hole per 8-hour shift is 
about 48 feet, or 1.2 inches per minute. 

Arizona Copper Co., Morenci, Ariz. Stoping with Waugh 
air-feed drills, in rather hard porphyritic ore, requiring no tim- 
bering and worked by the block caving system. Eight-hour 
shifts. In a period of 21 working days (Jan., 1910), 260 holes 
were drilled; total depth, 1,433 ft., average depth, 5J ft. Drill- 
ing time, 137 hours; delays, for picking down loose ground, etc., 
31 hours. Average drilling speed, excluding delays, 10.46 ft. 
per hour, or 2.09 inches per minute. 

Michigan Copper Mine, Rockland, Mich. Stoping in amygda- 
loidal brecciated vein matter, with Murphy air-feed drill, 628 
feet of hole were drilled at the average rate of 42 ft. per shift of 
9 hours, or 4.66 ft. per hour = 0.93 in. per min., including all de- 
lays. Two 6-foot holes were drilled on a test run at an average 
speed of 3 in. per minute. 

Snowstorm Mine, Idaho. Stoping in well-mineralized quartz- 



374 



COMPRESSED AIR PLANT 



ose ore, with Waugh air-feed drills, 14 holes were drilled as 
follows : 

Table XL VI. 



Hole 


Depth, 
Inches. 


Total Time, 
Minutes. 


Net Drilling 
Time, Minutes. 


Inches Drilled per 

Minute. 
















Total Time. 


Net Time. 


I 


15 


10 


8 


1.50 


1.87 


2 

3 


24 
8 


13 
19 


9 
10 


1.85 
hole lost. 


2.77 


4 

5 


25 
26 


15 
12 


8 

5 


2.16 


5.60 


6 


31 


19 


12 


1.82 


2.58 


7 


34 


20 


14 


1.70 


4.42 


8 
9 


44 

48 


II 
II 


7 
8 


4- 
436 


6.30 
6. 


10 
II 


11 


33 
17 


21 
10 


1.60 

2.82 


l& 


12 


51 


40 


15 


1.28 


3 40 . 


13 


46 


20 


12 


2.30 


3.82 


14 


49 


18 


II 


2.72 


4-45 


Total 


502 


258 


150 


30.08 


50.07 


Averages 


35-8 






2-15 


3-58 



South Crofty Mine, Cornwall, England. The makers of the 
Climax Imperial hammer drill state that in Feb., 1909, one 
of their machines, column-mounted, made 34 holes in granite,' 
averaging 37 inches deep, in 7 hours total time, or 3 inches 
per minute. 

Village Deep and New Reitfontein Est. Mines, South Africa. 
In Dec, 1908, and Jan., 1909, 203 holes, aggregating 669 feet, were 
drilled by a Climax Imperial drill in 10 J shifts, or 63.7 ft. per shift; 
a general average of 1.58 inches per minute. 

Barre Quarries, Vermont. The Sullivan Machinery Co. state 
that in Barre granite, their larger hand hammer drill, used in 
dimension stone work, makes a 6-inch hole in i to i J minute. 

Midlothian Colliery, near Richmond, Va. Sinking a rock slope 
in rather soft, coarse-grained sandstone, with three Hardsocg hand 



COMPRESSED AIR HAMMER DRILLS 



375 



hammer drills; 95 lbs. air pressure. Cross-section of slope, 7 ft. 
X 16 ft.; 27 four-foot holes = 108 linear feet per round; an average 
of about 18 rounds are made per month of 26 days, with three 
8-hour shifts, the corresponding advance of the slope averaging 
65 ft. Actual drilling time not recorded. 

Esmeralda Mine, Silverton, Colo. Stoping with Waugh drill; 
holes I and 2 in medium hard andesite, hole 3 in hard quartzose 
ore and hole 4 in rather soft vein rock: 

Table XL VII. 



Hole 


Depth, 
Inches. 


Total Time, 
Minutes. 


Net Drilling 
Time, Minutes. 


Inches Drilled per 
Minute. 




Total Time. 


Net Time. 


I 
2 
3 
4 


35 
58 
49 
62 


13 

27 

23 
14 


10 

21.5 
18 
12 


2.70 

2-15 
2.13 

4-43 


3-50 
2.70 

5.16 


Averages 


51 


19.2 


15 -4 


2.85 


3-52 



Burra-Burra Mine, Tennessee Copper Co. Stoping with 
Murphy and Waugh drills, in medium hard pyritic ore. Average 
speed for 9 "uppers," ranging from 52 to 72 inches deep and 
averaging 63.5 inches, was 2 inches per minute, total time, includ- 
ing changing bits. Maximum speed, 3 inches; minimum speed 
1.6 inches per minute. The usual duty per 9-hour shift is from 
40 to 50 feet of hole, or 4.4 to 5.5 ft. per hour = o.85 to i.io inches 
per minute, general average. 

Los Angeles Aqueduct, Los Angeles, Cal. Tunneling in medium 
to hard, close-grained granite, favorable for drilling; Leyner 
drills Nos. 7 and 9, mounted on a horizontal bar; air pressure 
about 90 lbs. Test runs were made to compare the work of the 
two sizes of machine.* 



* Abstracted from an article by J. B. Lippincott, Engineering News, April 22, 
1909, p. 449. 



i76 



compressed air plant 
Table XLVIIL 



No. 9 Drill. Weight of H.\mmer, 13^ Lbs. 600 Strokes per Minute. 



Hole. 


Depth, 
Inches. 


Total Time, 
Minutes, 


Net Drilling 
Time, Minutes. 


Inches Drilled per 
Minute. 


Total Time. 


Net Time. 


I 
2 
3 


79 
87 
96 


9- 
3625 
29.25 


7 

16.75 
1925 


8.77 
2.40 
3.28 


11.30 
5.20 
4.98 


Averages 


87-3 


24.83 


14-33 


4.82 


7.16 



No. 7 Drill. Weight of Hammer, 7 Lbs. 1,600 Strokes per Minute. 



I 

2 

3 


83-5 

83 

84 


7-75 
19-50 
19.50 


6.50 
16.00 
16.00 


6.08 

4-25 
430 


12.80 
5-19 
5-25 


Averages 


83-5 ^ 


15-58 


12.83 


4.88 


7-75 



The rock in which the first hole was drilled by each machine 
was much softer than that of the others. 

Field of Work. It may hardly be questioned that, for such 
work as the driving of tunnels or cross-cuts of large section, or 
for underhand stoping in wide veins and in general wherever deep 
holes can be advantageously drilled and blasted, the majority of 
the hammer drills (omitting large machines of the Leyner type), 
cannot be expected to compete with reciprocating drills. But, 
in connection with these operations, hammer drills may often be 
made useful as auxiliaries, not only for block-holing but for 
"squaring up," after the main rounds have been fired; that is, 
dressing the walls, taking up the bottom when the deep holes fail 
to break clean, etc. For other mining work, also, as for example, 
drifting and stoping, either breast or overhand, the air-feed 
machines are useful; and, when mounted on bars or columns, 
will often give quite as good results as the 2 or 2 J inch reciprocating 
drills, providing deep holes are not required. 



COMPRESSED AIR HAMMER DRILLS 377 

The comparison is most favorable to the hammer drill in 
quarry work — both dimension stone and block-holing — and when 
stoping in thin veins, especially where the values occur chiefly in 
narrow pay-streaks. In the latter case, shallow holes, with corre- 
spondingly light charges, are desirable, to avoid scattering the rich 
ore and breaking it with the poor ore or wall-rock, which would 
have to be sorted out. 

For shaft-sinking the case is not so clear. When the rock is 
solid with few fissures, slips, or bedding planes, reciprocating 
machines, capable of drilling 6, 8 or in large shafts even lo-foot 
holes, produce the best results. But, in shaly ground, or any 
rock that is full of slips and short fissures, deep holes often break 
imperfectly. It is here that the hammer drill can be used with 
advantage, because such ground is worked best by a relatively 
large number of shallow holes, and to drill them with reciprocating 
machines involves loss of time in shifting and setting up. 

General Conclusions. The opportunity for substantial varia- 
tions in the design of hammer drills is relatively limited, the largest 
differences being in the arrangement of the air ports and passages. 
Since both valve and rotating mechanism are usually omitted in 
the smaller sizes, most of these drills have but one moving part — 
the hammer — and their strong construction and simplicity adapt 
them well for the rough usage unavoidable in mine and quarry 
work. In the matter of repairs, therefore, they compare favorably 
with reciprocating drills. 

Economy in the use of air is also a feature of the hammer drill. 
The smaller machines use say 25 to 30 cu. ft. of free air per minute; 
the larger, including the air-feed drills (but omitting the "Water" 
Leyner), from 35 to 55 cu. ft. per minute. Air pressures of at 
least 80 or 90 lbs. (sometimes 100 lbs.) are generally recommended; 
that is, the same pressures as are suitable for the large reciprocating 
machines. Some foreign machines, for example, the Gordon, 
which has done remarkably good work,* seem to be best adapted 
for pressures of not over 50 or 60 lbs. 

* In the latter part of 1907, a series of tests was made on a number of small 
stoping drills, both reciprocating and hammer, at Johannesburg, South Africa. 



378 COMPRESSED AIR PLANT 

The objection is sometimes made to hammer drills that, in 
drilling dry holes, they raise much dust. With the small hand 
machine, the operator must necessarily stand close to the mouth 
of the hole and, when the latter is cleaned by exhausting through 
a hollow bit, the dust is blown back directly in the operator's face. 
This is specially troublesome — and hurtful — in drilling either 
breast -holes or uppers (above the horizontal). It can hardly be 
denied that hand hammer machines do not appear to best ad- 
vantage when drilling holes in such positions. The Murphy hand 
drill is provided with a ring of sponge, around the bit near the 
mouth of the hole. By keeping the sponge wet the dust difficulty 
is diminished, though not eliminated. Miners often neglect to 
care for such appliances. For work of the character mentioned, 
and for stoping in general, the air-feed machines, held by the opera- 
tor or mounted on a bar, are to be recommended, together with 
the use of water injected through the hollow bit, or an ex- 
ternal spray, as in the Climax drill, whenever the hole fails to 
clean itself. 

Solid bits have only a limited application for hammer drills. 
They can be used for shallow holes in the softer rocks, by watering 
the hole and frequently spooning out the sludge; or in dry rock, 
provided the holes are inclined at a sufficient upward angle to 
permit the cuttings to run out by gravity. But they are best used 
for cutting hitches for mine timbers, block-holing, and quarry work. 

An important feature of hammer drills with air-feed attach 
ment is, that they can be set up or shifted in less than two minutes, 
under normal conditions; while the time required for shifting an 

At the air pressure used (60 lbs. maximum), the Gordon drill stood first, drilling 
36 ft. 9 ins. in 4 hrs., in granite; the next best record being 33 ft. 2 ins., by the 
Chersen, a reciprocating machine. But it is probable that, with a higher air pressure 
the results of the test would have been quite different, and more favorable for some 
of the other machines, designed for higher pressures. Weaknesses of construction 
were subsequently discovered in the Gordon drill, which is not now on the market. 
Full details of these tests, with tabulations of the work done by each machine, were 
published in the proceedings of the Transvaal Institute of Mechanical Engineers, 
January 11, 1908. 

It may be added that since then the Climax hammer drill has done equally fast 
work. 



I 



COMPRESSED AIR HAMMER DRILLS 379 

equivalent 2 or 2j-inch reciprocating machine, mounted on a col- 
umn, is rarely less than 15 minutes and sometimes half an hour. 
The relative drilling capacity of the light hammer drills is thus in- 
creased. Still another saving in time results from the fact that 
the bit is loose in the chuck, so that for changing bits there are 
no chuck-bolts to be manipulated. 

Hammer drills supply a substantial need in mining and quarry- 
ing, and in their proper field of work will unquestionably be more 
and more widely used. 



CHAPTER XXII 

COAL-CUTTING MACHINERY 

Coal-cutting machines are extensively used to replace hand labor 
in the work of "under-cutting" the coal, preparatory to breaking 
it by blasting or wedging. The objects in view are: (i) To 
economize in the cost of mining; (2) to decrease the proportion 
of "fines" produced; (3) to increase the rate of production of 
coal from a given extent of mine workings. Coal cutters find 
their chief application in bituminous collieries, and under proper 
conditions it is unquestionable that by their use coal can be pro- 
duced more cheaply than by manual labor alone. They are 
designed to groove or undercut the face or breast of coal, close 
to the floor and to a depth of several feet. The mass so undercut 
is subsequently broken down by a few comparatively light blasts. 
Fig. 163 shows atypical case of a reciprocating or pick machine, 
working in a thin vein. 

Coal cutters may be divided into four classes : 

1. Endless chain machines. 

2. Rotary bar machines. 

3. Disc or circular saw machines. 

4. Reciprocating or pick machines. 

The last named imitates in some respects the operation of a 
miner's pick. All of them may be driven by either electricity or 
compressed air, though electricity is now generally employed for 
cutters of the first three types. These will be only briefly de- 
scribed. Pick machines are almost invariably operated by com- 
pressed air, no entirely satisfactory electric driven pick having 
yet been placed on the market. 

Endless Chain Cutters are built by several concerns, one ex- 
ample, the Jeffrey (type 16-D), being shown in Fig. 164. It has a 

380 



';! 



COAL-CUTTING MACHINERY 



381 



stationary bed-frame, consisting of two parallel steel channels, with 
the necessary cross-ties or braces. Within this frame is a T-shaped 
sliding frame, on the rear end of which is mounted the driving 
engine — a pair of small compressed air cylinders, say 5 in. diameter 
by 5 J in. stroke. The sliding frame carries an endless sprocket- 
chain, driven by a sprocket-wheel and gearing from the engine. 
Each alternate link of the chain is provided with a socket, in which 
is set a cutting tooth or bit. The head of the sliding frame is 




Fig. 163. — Sullivan Coal Pick, Working in a Thin Vein. 



composed of two parallel steel plates, as shown, between which, at 
the forward corners, are placed the idler sprockets carrying the 
chain. Mounted on each end of the frame are screw-jacks, for 
bracing the machine firmly in position while at work. The front 
jack is set against the face of coal, the other against a post or prop. 
The cutting bits, held in their sockets by set screws, are 
straight-edged and point forward in the direction of motion of 
the chain. They are so shaped and staggered with respect to one 
another, as to "cover" the chain and cutter head, and make a 
groove in the coal about 4 inches in height, or sufficient to permit 



384 COMPRESSED AIR PLANT 

the cutter head to enter freely as the work advances. The 
sliding frame is fed forward automatically as explained below. 

Fig. 165 shows the general plan and elevation of this machine, 
with some of the principal dimensions, and Fig. 166 is an enlarged 
plan and elevation of the compressed air engines and their acces- 
sory mechanism. Referring to Fig. 166 the engine crank-shaft a 
drives the countershaft h, through the gear wheels c and d. Keyed 
on h are: (i) The bevel gear e, engaging with bevel gear/, on 
the short vertical shaft g, which carries on its lower end the sprocket 
h, for driving the cutter chain; (2) the worm i, which, through 
the shaft j (and by another worm, as shown in the elevation), 
drives the gear wheel k, mounted loosely on the transverse shaft /. 
At each end of / are the small pinions m, m, engaging with long 
feed racks, bolted to the side channels of the stationary or main 
frame; (3) the small bevel gears n and 0, which, through the 
shaft p, drive the worm q, engaging with the gear r, also 
mounted loosely on the transverse shaft /. 

The effect of these two trains of gearing is to cause k and r to 
rotate in opposite directions, r having the more rapid movement. 
By means of the sliding clutch s on the shaft /, either ^ or f can be 
thrown in gear. When k is in gear the sliding frame carrying the 
driving mechanism, with the cutter head and chain, is fed slowly 
forward as the undercut advances. When r is thrown in gear, 
after the cut is finished, the feed js reversed and the sliding frame 
and cutter head are rapidly run back, preparatory to shifting the 
machine for the next cut. 

The depth of undercut made by chain machines ranges from 4 
to 7 feet; width of groove, from 39 to 44 inches. By shifting the 
machine, successive cuts are made side by side, over the required 
length of face. In regular operation, from 100 to 150 square 
yards can be undercut in 10 hours, depending on the character 
and hardness of the coal and upon whether the work is in "rooms" 
or "long-wall" mining. Two men are required: a machine 
runner, and a helper to remove the debris from the undercut and 
assist in handling the machine. Tests show that the power re- 
quired varies from 8 to 14 horse-power, an average of 12 horse- 




W/M^//, 




Y^^///^ 




^ 


\ 




\ 


t 




1 


rv«^ 




^^^>^ii^ 



lUii I i liJ 




COAL-CUTTING MACHINERY 385 

power being obtained from numerous tests in coals of varying 
hardness and quality. 

Chain cutters of a number of different makes are in use, oper- 
ated by either electricity or compressed air. Among them are 
those of the Jeffrey Manufacturing Co., Sullivan Machinery Co., 
Link Belt Machinery Co., and the General Electric Co. (Sperry 
machine). For long-wall mining some of them are self-propelling 
and operate continuously along a face of coal of any desired extent. 
A post is set at one end of the face and to it is attached a chain 
leading to the cutter, the chain being wound in by a small drum, 
geared to the driving engine. 

Rotary Bar Cutters are not so widely used as the chain machines. 
They have a bed-frame similar to that of the chain cutter and the 
inner sliding frame is fed forward in the same way. The cutter 
head carries a horizontal shaft or bar, in which is set a series of 
bits. This cutter bar is rotated by a sprocket-chain driven from 
the air engine (or electric motor) in the rear. The bar is from 
3 to 3J feet long, the width of undercut being slightly greater, 
owing to the projection of the end bits; speed of rotation of the 
bar, about 200 revolutions per minute. Under favorable conditions 
the time required to make a cut 5 to 5 J feet deep ranges from 6 
to 7 minutes. In room work, from 70 to 100 square yards can 
be undercut in 10 hours; more in long- wall work. The power 
required usually ranges from 14 to 16 horse-powxr. 

Disc or Circular Saw Cutters are made and used almost ex- 
clusively for "long-wall" mining. In the past they have been 
employed more commonly in Europe than in this country. Their 
general construction will be understood by reference to the ac- 
companying illustrations of the Jeffrey (x\merican) machine, style 
22-C. Figs. 167 and 168 are perspective views of opposite sides 
and Fig. 169 a plan and side elevation, on which only a part of 
the periphery of the cutter wheel is indicated. 

The driving machinery is compactly arranged on a bed-frame, 
mounted on wheels, to run on a temporary track laid along the 
face of coal. (The machine shown in the cut requires but a 
-single line of rails.) A cutter wheel, from 3 to 4^ ft. in diameter, 




u 



Q 



388 COMPRESSED AIR PLANT 

is supported by a heavy bracket on one side of the frame, and 
projects a distance nearly equal to its own diameter. On the 
periphery of the wheel is set a series of staggered bits, cutting a 
groove of sufficient height or w^idth, to admit the disc freely. 
The cutter wheel is driven by double reduction gearing, from a pair 
of compressed air cylinders, with quartering cranks. The speed 
of the wheel ranges usually from 15 to 30 revolutions per minute, 
depending on the character of the coal. When in operation the 
whole machine is automatically pulled along the face of coal by a 
small wire rope, made fast to a post at the end of the face, and 
wound in by a drum geared to the driving engine. 

The disc cutter is a useful machine for long-wall work, specially 
in thin or steeply pitching veins, where it would be difficult or im- 
possible to operate cutters of the other types. It works back and 
forth along the face, almost as well up the dip as down. The 
linear speed of feed is from, say, 8 to 25 inches per minute. These 
machines will undercut from 90 to 100 square yards per 10 hours, 
under average conditions. 

Among the English machines of this class may be mentioned 
the Gillott and Copley, Yorkshire, Rigg and Meiklejohn and the 
Winstanley. Though they vary in details, their general construction 
and operation are essentially as described above. 

Reciprocating or Pick Machines. — At the present time these 
constitute the most important of the four classes of coal cutters. 
In their general construction they possess many points in common 
with the reciprocating rock drills; which, in fact, have furnished 
the basis of the design of several well-known makes of pick ma- 
chine. All work without rotation of the piston and bit, since in 
undercutting there is no question of preventing rifling of the hole, 
as in rock-drilling. Variations in the valve-motion are noticeable, 
some of the designs being entirely different from the spool- and 
tappet-valve motions. 

The first pick machine was invented in 1858, by E. Simpkins, of 
Allegheny City, Pa., but it was of crude design, imitating the 
movements of the miner's pick. Next in order of seniority is 
probably the Harrison machine, invented in 1877. This was fol- 



COAL-CUTTING MACHINERY 389 

lowed in 1881 by the Yoch coal pick. From time to time both of 
these machines have been altered and improved in many of their 
features. Of later date are the Sergeant and the Sullivan picks. 
All of the above are operated by compressed air.* 

The general lines of this class of coal cutter, together with the 
mode of operation, are shown in Figs. 163, 17c and 171. The ma- 
chine is mounted on a pair of v/heels and when at work is placed 
on a wooden platform, about 3 ft. wide by 8 ft. long, w^hich slopes 
towards the face of coal, at an angle of, say, 5 degrees. By this 
means, the recoil of the blows is nearly neutralized by gravity 
and the machine is kept up to its work. The operator chocks 
the wheels with wooden blocks, sometimes strapped to his boots, 
and directs the blows by swinging the machine from side to side, 
with the supporting wheels as a fulcrum. As shown by the sev- 
eral cuts, the front cylinder head and piston rod are very long, 
to give the machine a sufficient reach. A horizontal width of 4 
or 5 ft. of undercut is thus readily commanded. The depth of 
cut is rarely greater than 5 feet. A helper clears away the debris 
with a light, long-handle shovel, and assists in moving and setting 
up the machine. 

Most pick machines run at speeds of 200 to 250 strokes per 
minute. The lowxr speed machines probably have some advan- 
tage, because, as each individual blow is directed by the operator, 
he can increase the efficiency of the work if he has time between 
strokes to point the pick in such a manner that it will do most 
execution. In coal of average quality, an undercut of say 4 ft. 
by 4 ft. in horizontal area can be made in 16 to 18 minutes. The 
platform can be shifted sidewise to the next position and the bit 
changed, if necessary, in 8 to 10 minutes. The height of undercut 
is 12 to 14 inches at the face, tapering to 3 in. or 3 J in. at the 
bottom. On completing the cut, the coal is shot down by light 
blasts; or, in gassy mines, is sometimes broken by wedging. 
Under favorable conditions, good operators can undercut, per 
shift, from 75 to 85 linear feet of face, to a depth of 4 to 4^ feet; 

* It may be added that an electric-driven coal pick, the Thomson-Houston 
solenoid, has been put on the market. 




pLn 





V v! 


i 


1^ 


;-ii^' 


<^ gSf- "^ ' ^ 


l^-- 


■:k 


v-- 




L 


■"s^. 


J^^ 


^p^ 


^ 


P 



Fig. 173. — Rotary Engine for Operating Valve of Harrison Coal Pick. 
Top View. 




Fig. 1 74. — Rotary Engine for Operating Valve of Harrison Coal Pick. 
Bottom View. 



394 



COMPRESSED AIR PLANT 



in competition trials much faster work is often done. Fair, average 
work would be from 60 to 65 ft. of undercut, 4 ft. deep, per shift. 

Harrison Pick Machine. — ^Fig. 172 shows a longitudinal section 
of models "P G" and 'T W" of this machine. The valve is a 
long, double spool, actuated through a crank and connecting rod by 
a small horizontal rotary engine set above the middle of the valve 
chest. A separate plan of the rotary, with its ports, is also shown 
in the cut; the main cylinder has double ports at each end, to 




Fig. 175. — Ingersoll-Rand Coal Pick. 



obtain cushioning of the stroke, and to enable the machine to 
run with a short stroke when desired. Figs. 173 and 174 are 
perspectives respectively of top and bottom of the valve motor for 
models W of the Harrison coal pick, the latter showing also the 
long slide valve and its connecting rod. 

These machines are made in heavy and light patterns, the 
former weighing about 700 lbs., the latter 500 lbs. The heavier 
pick will cut to a depth of 5 ft. and is specially adapted for " shear- 
ing"; that is, making vertical cut or grooves, on one or both sides, 
in driving entries for developing a coal seam, or for haulage- and 



COAL-CUTTING MACHINERY 



395 



air-ways. For this purpose, 
supporting wheels of 34 ins. or 
40 ins. diameter are provided, 
in order to raise the machine 
high enough to give the requis- 
ite reach. Smaller wheels are 
used for ordinary undercutting. 

" New Ingersoll " Pick Ma- 
chine (Fig. 175). — ^The valve- 
motion of this machine is 
developed from the spool-valve 
rock drills of the IngersoU- 
Rand Co., with the important 
difference, however, that the 
air ports are controlled by a 
combination of slide- and spool- 
or plunger-valves. Fig. 176 is 
a longitudinal section and 
Fig. 177 a diagrammatic sketch 
of the valves, ports and chest, 
developed in one plane to show 
the relations between the sev- 
eral ports and the operation of 
the valve motion. 

The main ports S, S^ of 
the cylinder O are controlled 
by the slide-valve G, on the 
back of which is a lug H, en- 
gaging with the double spool- 
valve F. Air is admitted al- 
ternately to the opposite ends 
of F through the auxiliary ports 
/, J^ which in turn are con- 
trolled by the auxiliary slide- 
valve K and its actuating 
spool-valve F^. This valve is 



^ 



\- 



Ph 



396 



COMPRESSED AIR PLANT 



thrown by air passing through the small ports A^, N^, N^ 
and A^^, which, by a cross-over arrangement as shown, connect 
on either side with the main ports S, S^. Thus the rear end 
of the chest of spool-valve F^ is connected with main forward 
port 5* and the forward end of F^ with the rear main port 5^; so 
that, when air is admitted to the main cylinder O through the rear 



Stop for governor valve 

D V Governor valve 




Fig. 177. — Ingersoll-Rand Coal Pick. Diagram of Valves and Ports. 



port S\ a small portion of it passes through N, N^ and drives back 
the valves F^ and K. This movement admits air from F^ through 
port J^ to spool-valve F and throws back the main slide valve G; 
thus opening main port S to live air and S^ to the exhaust. Some 
of the air passes from 5 through the rear auxiliary port N\ N^ 
to the valves F^ and K, by which first K and then G are thrown 
forward, reversing the main ports and completing the cycle of 
movements. As the main port S^ is of larger area than the port 
S, the forward stroke of the machine is more powerful than the 
back stroke. 



COAL-CUTTING MACHINERY 397 

In addition to the above parts there are two small regulating 
screws L and D, by which the operator can regulate the speed 
of the auxiliary valves F^ and K and hence of F and G. But, 
while the effect of the regulating screws is to govern the speed of 
stroke, they do not greatly influence the force of the blow. The 
forward stroke is cushioned, as follows: When the piston passes 
the double port P the exhaust ceases, since at this point the bit 
should strike its blow. Hence, if the pick misses the coal or should 
momentarily be swung out of contact with it, the piston on advanc- 
ing beyond P compresses the air in the forward end of the cylinder 
to a pressure higher than that in the valve chest A'^, which is con- 
nected with the cylinder by the small port W. This high pressure 
air forces back the governor valve F, wholly or partly cutting off 
the inlet air on its way to valve K, so that the machine will run at 
reduced speed and force until the bit again strikes the coal before 
the piston has covered the port P. The regular exhaust then takes 
place, relieving the pressure transmitted through W to the governor 
valve, which opens automatically and the machine at once resumes 
regular operation. By means of the stop D the throw of the 
governor valve may be adjusted as desired. 

The air on which the piston cushions is confined by the spring 
check-valve, T, in the port 5; and the back stroke is begun by the 
elasticity of the cushioned air. No live air can enter this end of 
the cylinder until the piston has advanced far enough on the back 
stroke to reduce the pressure in the cylinder below that of the live 
air in port 5, plus the resistance of the check valve spring. A 
smaller degree of cushioning is similarly produced on the back 
stroke by the arrangement of the port S^ and valve 5^ 

The Ingersoll pick is made in three sizes of cylinder, viz., 4 J 
in., 5 in. and 6 in. diameter, but there are five models, varying 
in weight, power and length, according to the depth of undercut 
desired. The maximum depth of cut ranges from 4 to 6 feet and 
the gross weights of machine, from 550 to 950 lbs. Standard 
wheels for mounting the pick are 14 in. and 1 7 in. diameter. Larger 
wheels may be used for shearing. 

Sullivan Pick. — Fig. 163 is a general view of the machine. It is 



398 



COMPRESSED AIR PLANT 



made in three sizes, with cylinder bores of 4^ in., 4J in. and 5^^ 
in., which undercut to maximum depths of from 4 J to 6 feet. The 
weights range from 650 to 850 lbs. Air consumption of the 
4i-inch machine is about 115 cu. ft. of free air per minute; of the 
5j-inch machine, 135 cu. ft. 

A longitudinal section of one of the larger sizes is given in 
Fig. 178, accompanied by a descriptive list of the principal parts. It 
will be seen that in some of its features this machine differs greatly 
from those previously described. The spool- or piston-valve 
(123), by its reciprocations, throws the long fiat valve (126), which 
in turn controls the main ports (148 and 149) and the secondary 
ports between them. These secondary ports serve to produce 
cushioning at each end of the stroke, in a manner similar to that 
of the Ingersoll-Rand machine. A check- valve (130) is inserted 
in the forward main port (149), and, when the pick does not strike 
the coal, the piston runs forward far enough to form an air cushion 
in the front end of the cylinder. This closes the check-valve, 
and prevents immediate admission of live air on the return stroke, 
the first part of which is made by the cushion air. Injury to 
the cylinder head is thus prevented, as well as annoying shocks 
to the machine runner. 



List of Principal Parts (Fig. 178). 



100. 


Piston (bare). 


121. 


Ring for 122. 


104. 


Rifle Nut. 


122. 


Packing leather (small) for 123. 


105. 


Rifle bar, with gear. 


123. 


Valve (piston). 


106. 


Seat for ioq. 


124. 


Buffer for 123. 


107. 


Spring pointer for io8. 


126. 


Valve (flat). 


108. 


Stem for adjusting io6 


130. 


Check-valve. 


109. 


Reverse valve. 


134- 


Plug for oil hole. 


no. 


Valve plate. 


135- 


Pick. 


III. 


Cover for no. 


136. 


Chuck. 


113- 


Spring for 115. 


137- 


Head (front) for 142. 


114. 


Regulating valve. 


139- 


Bushing in 137. 


115- 


Index lever for 114. 


140. 


Packing leather for 100. 


116. 


Head (bare) for 127. 


141. 


Collar for 140. 


117. 


Packing leather (large) for 123. 


142. 


Cyhnder (bare). 


118. 


Ring for 117. 


144. 


Trunnion for wheel. 


120. 


Binding screw for 118 and 122. 


147- 


Drift-Key for backing out pick. 



COAL-CUTTING MACHINERY 



399 




a 
o 



400 



COMPRESSED AIR PLANT 



In the hollo\^' piston is set a rifled-nut (104), with which engages 
the rifle-bar (105). As the forward end of the piston rod has 
flattened sides, making it nearly square in section, it cannot rotate 
like that of a machine -drill. Hence, at each stroke the rifle-bar is 
caused to rotate. The small gear cut on the back end of the rifle- 
bar causes the rotation of the reverse valve (106 and 109). This 
valve in turn controls the movements of the spool- valve (123), 




Fig. 179. — Ingersoll-Rand "Radialaxe" Coal Cutter. 

through the ports shown by dotted lines in the rear end of the 
cylinder. By means of a regulating valve (114), on the back of 
the main valve chest, the operator adjusts the speed of stroke, as 
the conditions under which the machine is working may require. 
The Sullivan pick runs at a moderate speed, which has the ad- 
vantage of enabling the operator with more certainty to direct each 
blow where it will do the most efficient work. 

Ingersoll-Rand *< Radialaxe " Coal Cutter.— This machine is 
intended for shearing in entry work, as well as for undercutting. 




Fig. i8o. — IngersoU-Rancl " Radialaxe " Coal Cutter. 



402 COMPRESSED AIR PLANT 

As originally designed it consists essentially of a long-stroke rock 
drill, mounted on a column, and provided with a worm and worm- 
wheel sector. A hand wheel on the worm spindle enables the 
operator to swing the entire machine while at work, in either a 
horizontal or vertical arc. Long-shank bits are used, to give the 
machine the necessary reach. 

The latest form of the '' Radialaxe" as shown inFigs. lypand i8o 
is an adaptation of the Temple-IngersoU Air-Electric drill. In this 
design the gearing for swinging the machine is changed in some 
unimportant details. The bit, or group of bits, as shown, is set 
in a rosette socket, held by friction only on the tapering end of the 
drill shank, which is similarly inserted in the deep socket of a long 
chuck. The individual bits are thus readily removed for sharpen- 
ing and replacement when broken. 

Pneumelectric Coal Puncher. — Amongst the coal picks this 
machine occupies a class by itself. As its name implies, it com- 
bines the use of compressed air and electricity in a single piece of 
mechanism, consisting of a small electric motor, which drives a 
pair of independent pistons in a cylinder. This design may have 
been suggested by the Temple rock drill. Figs. i8i and 182 show 
the general construction. The casing T contains the motor, with 
its armature shaft J. in a vertical position. To change its rotary 
motion to the rectilinear motion required for the piston, the 
following device is employed. The driving pinion B engages with 
the large horizontal gear-wheel C (Fig. 182), which has a solid 
web, carrying a stud D (Figs. 181 and 183). Mounted on D is a 
small gear £, and a crank with crank-pin G. Within the main 
gear C, and attached rigidly to it, is another gear F, with 
internal teeth, which engages with and drives the gear or crank- 
pinion E\ F having 66 and E 33 teeth. Referring to the two dia- 
grams in Fig. 183, it will be seen that the gears are so propor- 
tioned that the stud D revolves in a circle concentric with F and 
of one-half its pitch diameter. The gear E, which revolves freely 
on D, is carried with D, and causes the crankpin G and the cross- 
head H to reciprocate between the guides. To the cross-head 
is attached the piston rod /, with its piston J (Fig. 181). 



COAL-CUTTING MACHINERY 



403 




404 



COMPRESSED AIR PLANT 



The cylinder T' contains two pistons, entirely unconnected with 
each other: the rear or driving piston 7, already noted, and the 
forward piston K, with its rod L and chuck A^, for holding the 
bit Y. In starting the machine on its first forw^ard stroke, piston 
J simply pushes K forward, air meantime entering the cylinder 
behind the piston, through the valve O. On the back stroke of J, 
the air in the rear of the cylinder is compressed, occupying the 
clearance space and the passage between the cylinder and the valve 




(a) ih) 

Fig. 183. — Pneumelectric Coal Puncher. Diagram of Gearing. 



O. At the same time the air between the pistons is rarefied, thus 
causing K also to make its return stroke by suction, while air 
enters freely through a port at R. At the end of the back stroke, 
the air passages below the valve O, referred to above, lie between 
the pistons, whereby the charge of compressed air enters the 
cylinder and drives the piston K forward on its first regular stroke. 
The stroke is cushioned on air confined in the end of the cylinder, 
after K passes the port at R. Piston K, having completed its 
forward stroke, is followed by piston /, the air between them 
being discharged into the atmosphere through the valve S. The 
return stroke is then made by both pistons, as at first. 

The diameter of the cylinder is 6 J inches, and the clearance 
spaces at the rear end are proportioned to produce a final or work- 
ing pressure of 95 to 100 lbs. The motor is designed to run at 



COAL-CUTTING MACHINERY 



405 



three speeds, under the control of the operator, giving to the pick 
140, 160 or 180 strokes per minute.* It is stated that 7 H.P. are 
required to run the machine, which is in successful operation at a 
number of collieries. The behavior of the motor, while at work, 
is shown by the following records, taken after two hours' operation. 



Test. 


Input. 


Temperatures, Degrees Centigrade. 


Amps. 


Volts. 


Room. 


Armature 
Winding. 


Commutator. Brush. 


Field Coils. 


I 


26 

24 


220 
220 


17 
24 


79 
73 


75 
64 


77 


70 


2 


not recorded. 



The temperatures include room temperatures prevailing during 
the tests. It may be noted, respecting this machine, that, as the 
compressed air is released at the same point of every stroke, the 




Fig. 184. — Stanley Heading Machine for Collieries. 



power required of the motor is constant ; in other words, the motor 
is not subject to overloads. 

The Stanley Header, originally brought out in England, is in- 
tended for development work in collieries, driving circular head- 

* This description is abstracted in the main from an article by Timothy W. 
Spragtie, in Mines and Minerals, April, 1908, p. 427, 



4o6 COMPRESSE© AIR PLANT 

ings, for entries, airways, etc. In its earlier form (Fig. 184) a 
heavy cross-head, a, mounted on a central screw shaft, h, and 
carrying at each end a horizontal arm, c, cuts an annular groove 
from 3 to 4 inches wide. A central core of coal is thus left, which 
either breaks up and is shovelled back as the work advances, or is 
blasted or wedged down from time to time, if necessary. The 
operating mechanism consists of a pair of compressed air cylinders, 
about 9 ins. by 9 ins., from the crank-shaft, d, of which the screw- 
shaft is driven by gearing. Differential gearing is used to produce 
the feed, as shown. The whole is carried in a frame on wheels, 
held firmly when in operation by jack-screws set against the roof. 
The machine is narrow enough to permit a man to pass alongside 
to the front, to throw back the broken coal and keep the cutter 
head free. 

Several modifications of the Stanley Header have been intro- 
duced in this country and abroad. By one of them, the entire 
section of the heading is taken out in a single operation. To 
accomplish this, the cross-head and arms, on the central spindle, 
are replaced by a very fiat, cone-shaped head, which carries a 
number of individual cutter bits, arranged in diametral lines on the 
surface of the cone. The circular paths traversed by these bits 
cover one another, so that the whole mass of coal is broken up. 

This modification of the original machine has done good service 
in some of the bituminous mines of western Pennsylvania, such 
as those of the Frick Coal and Coke Co. The average speed of 
advance, under favorable conditions, is 2 J to 3 feet per hour, 
including the time occupied in moving and setting up after a run. 
In one case 2,254 linear feet of entry were driven at an average 
speed of 17 ft. per 9 hours, and an average working cost of 40 
cents per foot. In a mine of the Frick Co., 1,475 ^^^^ have been 
driven without a parallel entry for air connections, the exhaust air 
being discharged backward through an 8-inch pipe. This exhaust 
assists in ventilating long headings. 

For rapid development work, as in opening a bituminous col- 
liery for the "long-wall" method of mining, the Stanley Header 
is specially advantageous. It has been used, also, by the Col- 



COAL-CUTTING MACHINERY 407 

orado Fuel Co., with the following results, as compared with 
hand work: 

Hand Labor. 

2 men, i lo-hr. shift $4.00 

Paid to men for coal produced in driving, 4I tons @ 50c 2.25 

Cost $6.25 

Distance driven in 10 hours, 3 feet. 

Machine Work. 

I operator, $3.00; i helper, $2.50 SS-So 

3 shovellers to load coal behind machine @ $2.00 6.00 

Compressed air, repairs, depreciation and interest 3.50 

Squaring up corners, for timbering and track 5.00 

Cost $20.00 

Distance driven in 10 hours, 20 feet. 

Crediting to the machine work the coal produced, viz., 15 J tons 
@ 50c. loaded, the net cost for the 20 feet of entry was $12.25, 0^ 
$1.84 per yard. This shows a substantial gain over hand work, 
both in speed and cost. 

Auger Drills. — Though not strictly in place here, reference may 
be made to the rotary auger drills, for boring holes for blasting in 
coal, rock-salt and other soft material. They are operated by 
small compressed air engines — some also by electricity. 

Two of these are built by the Inger soil -Rand Company. The 
first is of the breast-drill type, similar in general form to a ma- 
chinist's breast drill. It has a 3-cylinder motor, which can readily 
be reversed for withdrawing the bit from the hole; total weight, 
exclusive of the bit, 18 lbs. The second, a heavier machine, is 
designed to be mounted, on a column or bar. 

A compressed air auger drill, mounted on single or double 
column, is made by the Jeffrey Manufacturing Co. Total weight, 
for a 6 foot vein, 183 lbs. The speed of the engine is about 3,000 
revolutions, and that of the threaded feed shaft, 850 revolutions 
per minute. 

Comparison of Coal Cutters. — ^The chain and disc machines 
work best in clear coal, of uniform quality and not too hard. 



4o8 COMPRESSED AIR PLANT 

For hard, "bony" coal, or coal containing streaks of pyrites, or 
"sulphur balls," the pick machines are preferable; because the 
operator can regulate the strength of the blow and so direct the 
machine as to cut around a hard place, when necessary. For this 
reason, however, the pick machine requires more skill on the part 
of the operator. 

Somewhat less slack and fines are made by the disc and chain 
machines, as the volume of undercut is smaller; but, taking into 
account the greater flexibility and variety of work possible with the 
picks, the latter are in many respects advantageous for all-around 
work. Moreover, in the numerous mines whose product goes 
chiefly to coke ovens, the larger quantity of fines made by the pick 
machines is immaterial. Also, in solid, hard coal, the higher 
undercut of the pick machines causes a more complete breaking up 
of the whole mass when blasted down, and the coal, therefore, is 
sometimes more readily loaded. 

For chain and disc cutters a fairly good roof is desirable; other- 
wise props may have to be set so close to the breast or long- 
wall face as to interfere with the manipulation and shifting of the 
machines. Coal cutters of all the types can be worked in seams 
as thin as about 30 inches, or even less; though they are more 
conveniently operated in seams not less than 36 to 42 inches 
thick. The continuous-feed chain and disc machines are specially 
useful for thin, pitching seams in long-wall work, as they will 
operate with almost equal facility either up or down the pitch. 

The "mining rate," or cost of mining by hand, together with 
the character of the coal seam, will usually determine whether 
coal cutters can be economically applied in a given mine or district. 
In general, it may be stated that in seams of average quality and 
thickness, when the cost of hand mining in the district is not less 
than 50 to 55 cents per ton (including loading the coal), a saving 
may be effected by introducing machines. 



CHAPTER XXIII 

CHANNELING MACHINES 

Originally, channeling machines were used almost exclusively 
for getting out dimension stones in quarry work. Of late years, 
however, they have been employed in increasing numbers for 
certain kinds of rock excavation, where it is desired to have 
smooth, uniform walls; for example, rock cuttings for railroads, 
canals and water-wheel pits for power plants. They are best 
adapted for cutting the softer rocks, like limestone, most of the 
sandstones, slate, shale, etc., though they may be used also for 
some of the varieties of granite, gneiss, porphyry, schist, and other 
metamorphic rocks. Hard rocks are best quarried by drilling 
rows of holes, with wedging or blasting. 

In a certain sense channelers resemble reciprocating rock drills, 
a single bit or a "gang" of bits being attached to the piston rod. 
But, instead of drilling a series of round holes, the channeler, as 
its name implies, cuts a continuous, narrow groove, without rota- 
tion of the piston and bit. In its typical form, the machine is 
solidly supported on a heavy carriage or truck, generally mounted 
on a track laid along the line to be cut. The motive power may 
be compressed air, electricity, or steam. By means of an auxiliary 
engine and worm gearing, the whole machine, while at work, is 
fed forward automatically at a suitable speed. The general con- 
struction of a standard compressed air-driven channeler will be 
understood by reference to Fig. 185. Another design, a track 
channeler for cutting marble, with adjustable mounting and a 
reheater mounted on the carriage, is shown in Fig. 186. 

General Construction. The construction of channeling ma- 
chines is varied to suit the conditions of work: i. For cutting 
vertical channels only, the rigid head machine is used; that is, the 
standard supporting the cylinder and accessories is non-adjustable, 

409 



412 



COMPRESSED AIR PLANT 



being permanently fixed in an upright position. Fig. 187 shows a 
steam-driven channeler of this class. It is employed for canal 
and railroad cuttings, general rock excavation and for quarrying 




Fig. 187. — Sullivan Rigid Back, Steam-Driven Channeler. 



where the strata are horizontal or nearly so. 2. For quarrying 
building stone lying in inclined stratified beds, like most lime- 
stones, the channels must generally be cut at right angles to the 



CHANNELING MACHINES 



413 



bedding planes; hence, the mounting of the cutting engine is 
adjustable, for making a channel at any desired angle to the 
vertical. The cylinder with its appurtenances is swivelled on 
its supporting frame or standard; or the frame may be provided 




Fig. 188. — Sullivan Adjustable Back Air-Driven Channeler. 

with T-slots (resembling those of the table of a planer), by means 
of which the cylinder is bolted firmly in the required position. 
The supporting frame, in turn (as in Fig. 186, of the Ingersoll- 
Rand Ram Track Channeler), may be swung back to a nearly 
horizontal position, for making angle wall cuts along a quarry 



414 COMPRESSED AIR PLANT 

face. In different designs, the minimum swing-back angle varies 
from 15° or 20° to 33° to the horizontal; but it is not often 
necessary to cut with these machines at less than 45°. Fig. 188 
shows a Sullivan channeler of this class. 3. A third form, 
designed specially for making horizontal channels, or " under- 
cuts," is used less frequently than the others. An Ingersoll- 
Rand machine of this type is shown in Fig. 189. As indicated 
in the cut, the head may be bolted to either end of the 
carriage. 4. Lastly, a light machine, which is in effect a large 
rock drill, may be mounted on a ''quarry bar" — a long, hollow 
bar, supported at each end by a pair of inclined legs.* This is 
generally used for drilling a row of holes placed close together, 
the partitions between which are afterward cut out by a "broach- 
ing" bit. Within a few years, a modification of the quarry-bar ma- 
chine has been brought out by the Ingersoll-Rand Company. 
It is a true channeler, mounted on a heavy swivel plate, which 
slides on a pair of horizontal bars, about ten ft. long, supported by 
inclined legs (Fig. 190). The whole machine is fed along the 
bars automatically by a small, 3-cylinder engine, which actuates 
a traveling feed-nut, engaging with a threaded shaft between the 
bars. 

There are many variations in construction in the above- 
mentioned classes of channeler, to adapt them to local conditions. 
Among other machines, of entirely different design, may be men- 
tioned the Wardwell and the Bryant channelers. The Wardwell, 
a heavy machine operated by steam only, has been in successful 
use for many years. It is intended for making vertical channels, 
a gang of bits being set in a massive frame, which is given an up 
and down movement, something after the manner of a jumper 
drill. 

For the first three classes of channeler a gang of from three to 
five bits is employed. These have long square shanks and are 
set closely side by side, the cutting edges being alternately at right 
angles and at 45° to the direction of the channel. This arrange- 

* In one of the Sullivan models, the legs are replaced by vertical standards, 
each carried on a small wheeled truck. 




H 



CHANNELING MACHINES 417 

ment forms practically a succession of Z-shaped bits, and insures 
the cutting of a regular channel with smooth walls. The bits 
are clamped firmly in a heavy chuck, attached to the piston rod 
of the engine, and are guided either by a cross-head or (as in some 
of the Ingersoll-Rand patterns) by a pair of roller guides. Some 
saving in power has been realized by the introduction of the roller 
guides; they eliminate part of the weight due to the cross-head, 
which must be lifted at each stroke, and the friction loss is reduced. 
The Sullivan Company builds a duplex or double-head machine. 
There are two cylinders, side by side on a heavy frame, each with 
its gang of bits and operated by a single valve-chest. As the blows 
alternate, one piston making its down stroke, while the other is on 
the up stroke, the machine can be run at high speed without exces- 
sive vibration. Its working capacity is correspondingly greater 
than that of the single-cylinder machines. When the plant con- 
sists of a few machines only, they may be advantageously driven 
by steam (Fig. 187) ; but, for large-scale work, a higher degree of 
economy results from the employment of compressed air, furnished 
by a central plant. Each machine is then provided with its own 
reheat er,* mounted on the carriage (Fig. 186). Air pressures 
generally range from 85 to 1 10 lbs. 

While at work the main cylinder of the channeler is raised 
and lowered in its guide shell by a screw-feed, operated auto- 
matically or by hand. The hand feed is rarely employed except 
for the smaller machines. The automatic feed may be caused 
either by an independent engine, similar to that used for the 
longitudinal feed of the quarry -bar machine, already referred to; 
or, by a chain and sprocket drive, from the engine which 
furnishes the propelling power along the track. The chain feed, 
as used in the Sullivan channelers, is shown in Figs. 185 and 187. 
Most of the Ingersoll-Rand channelers are provided with the 
independent feed engine, which is of the 3-cylinder type, very 
small and compact in design. In either case, when the cut has 
reached the required depth, the feed is reversed and the entire 

* See Chapter XIX, on Reheaters. 



41 8 COMPRESSED AIR PLANT 

head, with its accompanying parts, is raised preparatory to making 
the next cut. 

Depth of Cut and Speed of Work. The heaviest channelers — 
those with rigid back or standard — will cut to depths of from 8 
to 15 or 16 ft., according to the character of the stone; the swing- 
back and bar machines will cut from say 6 to 10 or 12 ft., and 
undercutting machines up to 7 ft. For starting a channel, the 
width of a bit is from ij to a maximum of 4 inches, depending on 
the depth of cut to be made and on the nature of the stone. 
The gauges of the successive bits are generally reduced by ^ 
inch each, the finishing bits usually cutting a width of i J inch. 

The cutting capacity of channelers varies greatly. It is largest 
in the softer stones, when of uniform texture and quality, and in 
fully developed quarries, where the work is systematic and the 
stone lies below the zone of weathering and surface disintegration. 
In sandstone of average hardness and under favorable conditions, 
from 250 to 300 sq. ft. of channel may be cut per 10 hours by the 
heavy machines; or, including all stoppages and delays, from 4,000 
to 4,500 sq. ft. per month; in the softer sandstones and limestones 
higher duties are obtainable. The swivel-head and other adjust- 
able channelers are lighter than the fixed-back machines and in 
the same kind of stones their rate of work is generally slower. 
Machines working in rather hard marbles, like those of Rutland, 
Vt., will cut from 2,300 to 2,500 sq. ft. per month, or an average 
of 85 to 100 sq. ft. per day. A single day's work, however, will 
often greatly exceed these figures. In hard marble or limestone, 
the smaller bar machines will cut an average of say 40 sq. ft. per 
10 hours and up to 125 sq. ft. in softer stones. For hard gneiss, 
or schist, like that of New York island, an average duty would be 
65 to 70 ft. per day. 

Tables XLIX and L, showing dimensions, weights, and other 
data, of the channelers of two well-known builders, will further 
illustrate the features of these machines. 

Recently, the Ingersoll-Rand Company have applied the prin- 
ciple of their "Air-Electric" rock-drill in the design of the cylinder 
and air compressing mechanism of a track channeler (Fig. 191). 



CHANNELING MACHINES 



419 



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422 COMPRESSED AIR PLANT 

That is, an electric motor, mounted on a carriage, drives a single- 
acting air compressing pulsator, which is connected to the channeler 
cylinder in a manner similar to that of the Temple-Ingersoll rock- 
drill, described in Chapter XX. For each double stroke of the 
pulsator, there is a blow and return stroke of the channeler piston; 
the speed of stroke being thus controlled by the speed of the electric 
motor which furnishes the power. Favorable results, both as to 
power cost and maintenance, have already been secured. This 
channeler has a swivel head and a swing-back support, and is 
therefore suitable for varied quarry service. 



I 



CHAPTER XXIV 
OPERATION OF MINE PUMPS BY COMPRESSED AIR 

It is intended here to deal only with that part of the extensive 
subject of mine drainage which has to do with the employment 
of compressed air as a motive power. Under this head there are 
three general forms of apparatus: 

1. Direct-acting pumps, single-cylinder, duplex, or compound. 

2. The air-lift pump. 

3. Pneumatic displacement pumps. 

In this chapter the first class only will be considered. 

Simple, Direct-acting Pumps. Notwithstanding the general 
similarity in the behavior of steam and compressed air, when used 
in the cylinders of direct-acting pumps there are some important 
points of difference. By first considering briefly the construction 
of the types of pump in common use the results obtainable from 
the employment of compressed air can best be set forth. 

The development of the direct-acting pump dates from Henry 
R. Worthington's invention in 1841; and the greater part of all 
the pumping in the mines of this country, and much of it in other 
countries also, is done by pumps of this class. The cylinders are 
set tandem, the power being transmitted from the steam to the 
water cylinder through a piston-rod common to both. As there 
are no rotating parts, the length of stroke is controlled by the ad- 
mission and exhaust of the steam. In all the simple pumps the 
valve motion involves the use of an auxiliary valve, whose move- 
ments are governed by the reciprocating movement of the piston, 
and which in turn operates the main valve. The duplex form 
consists essentially of two simple pumps, set side by side, with an 
inter-dependent valve motion ; that is, the valve of each is ope- 
rated positively, through a system of levers, by the movement of 
the piston of the other side. 

423 



424 COMPRESSED AIR PLANT 

Though direct-acting pumps are strong and reliable, simple 
in construction, and occupy but little space, they are extremely 
uneconomical machines, unless the steam cylinders are com- 
pounded. It is hardly necessary to say that this ought not to be 
the case. Pumping is an operation that should be conducted 
economically, especially in connection with mining, where the 
pumping of water is classed as "dead work." Moreover, the 
conditions in themselves are not unfavorable. A pump works 
under a practically constant load, from the beginning to the end 
of each stroke, the only necessary variation — which need not be 
large — occurring at the instant the discharge valves open. 

The trouble is that, in attaining compactness, simplicity, and 
moderate first cost, the power is not applied in simple, direct- 
acting pumps to the best advantage. As there is a constant 
load, but no fly-wheel to equalize the power, steam must be ad- 
mitted at full pressure throughout the entire stroke; otherwise 
the piston would be unable to reverse, and would come to a 
standstill. Such a pump must work practically without cut-off, 
and therefore a cylinderful of steam, nearly at initial pressure, is 
exhausted at each stroke. In some pumps the terminal pressure 
is quite as high as the initial. A duplex, non-compound pump, 
having a positive valve motion, may at times be even a more 
extravagant steam-consumer than a single-cylinder pump, since 
one piston may reach the end of its stroke before the other is 
ready to reverse its valve. In such case the momentum of the 
incoming steam fills the cylinder at initial pressure at the moment 
of exhaust. 

For steam-driven pumps there are several ways of improving 
these conditions : 

I. The adoption of compound or triple expansion cylinders. 
This type is suitable for the larger sizes of pump, and its use is 
increasing for mines whose depth and quantity of water warrant 
the higher first cost. The space occupied is but little greater than 
for simple pumps of the same capacity, and satisfactory results are 
obtained when they work under proper conditions and with 
sufficient initial pressure. 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 425 

2. While retaining the tandem form, a fly-wheel may be intro- 
duced, driven from the cross-head or from the steam-cylinder 
connecting-rod. This is a reversion to a type of pump long ago 
discarded for general service in this country, in favor of the 
simpler but less efficient form with no rotating parts. Although 
such a pump occupies niuch more space and its first cost is in- 
creased, there can be no doubt as to the advantages of being able 
to use the steam expansively, without the necessity of compound- 
ing. A large number of pumps of this description are now em- 
ployed in mines; many of the Riedler pattern and some of less 
elaborate and expensive design, such as the Prescott and others, 
in which an early cut-off — at one-quarter or even one-eighth 
stroke — is satisfactorily adopted. 

Notwithstanding the advances made along these lines in the 
mechanical engineering of pumps and the added economy gained 
in their operation, it has been very generally assumed in the past 
that similar economies are not attainable when compressed air 
instead of steam is employed as the motive power. Yet the ad- 
vantages accruing from the utilization of compressed-air trans- 
mission in mines are marked. As the heavy losses due to radia- 
tion and the condensation of steam in pipe-lines are avoided, the 
transmission of power by compressed air may be conducted with 
a high degree of efficiency. No difficulty exists as to the disposal 
of exhaust steam underground, nor is any danger to be appre- 
hended from the rupture of a compressed-air pipe, while the 
bursting of a steam pipe in a shaft or in the mine workings may 
cause serious trouble. The failure to realize these advantages, 
and the unsatisfactory results obtained in most cases from com- 
pressed-air-driven pumps, are due largely to the fundamental 
differences in the behavior of steam and compressed air when 
used in a motor cylinder. In Chapter XVII reference has been 
made to the reduction of cylinder temperature accompanying 
the expansion of compressed air. The point of cut-off being the 
same, this causes lower terminal and mean pressures with air 
than with steam. In other words, at a given initial pressure and 
without reheating, a cylinderful of air develops less power. 



426 COMPRESSED AIR PLANT 

This property of air, together with the fact that it does not 
condense, indicates clearly that steam and compressed air are not 
equally well adapted for use in an engine of the same design. It 
is not easy to understand, therefore, why mechanical engineers 
and especially pump-builders have not given more attention to 
the production of pumps properly designed for the use of com- 
pressed air. Few, if any, other branches of motor-engine prac- 
tice have been so neglected. Lack of information among users 
of compressed air is responsible in part ; in addition to which it is 
not generally realized that relatively unimportant modifications, 
at small cost, would produce much better results. Users of the 
ordinary steam pump have become accustomed to its low econ- 
omy, and, because it is strong and serviceable, it is apt to be ac- 
cepted without question when compressed air is used instead of 
steam. But in applying compressed air to the inefficient single- 
cylinder pump, as usually designed for steam, the net result is no 
better, and may be even worse, than that obtained from, steam. 
The clearance spaces are large and, as the air is admitted to the 
cylinder throughout full stroke, it is used in a w^asteful manner. 
Moreover, the stroke is often shortened by imperfections in the 
valve action. 

Another unfavorable feature of mine pumps driven by com- 
pressed air is the frequently improper selection of the cylinder 
proportions and arrangement of the plant. In mines having a 
number of levels the pumps are distributed according to varying 
requirements as to height of lift and quantity of water to be raised. 
The lowermost pump may have to work under a heavy head; 
others under a head of only 100 or 200 feet. As all are usually 
operated from the same pipe line and under a common air press- 
ure, it is clear that the dissimilarity of working conditions must 
be met by proportioning the water and power ends of each pump 
according to the work to be done. But, through error or care- 
lessness, the power end is often badly out of proportion, the tend- 
ency being to err on the side of furnishing too much power. 
The steam (or air) cylinder may be of such size as to require a 
pressure of only 30 or 40 lbs. per sq. in., while the pipe-line press- 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 427 

ure is 70 or 80 lbs., as usual with mine compressor plants. So it 
often happens that the deepest pump in the mine is the only one 
operating under a proper pressure. The cylinders of the others, 
even if running under throttle, are filled with air at full pressure 
when exhaust takes place.* 

The difficulty with common direct-acting pumps is thus two- 
fold : the air is used without expansion, and the pressure is often 
higher than is necessary. Recognizing, however, the convenience 
with which the inexpensive, ready-made single-cylinder pumps 
may be installed, and that in many cases efficiency of operation is 
really a secondary consideration, a few points will here be dis- 
cussed as to their employment, and the volume of air required 
for a given quantity of work. Questions relating to the 
expansive use of compressed air for pumps will be taken up 
afterward. 

Cylinder Dimensions of Simple Pumps. In calculating 
the sizes of cylinders for a simple, or single-cylinder pump, to 
work under given conditions, the dimensions of the water cylin- 
der must first be determined. There are three variables to be 
dealt with, viz: diameter, length of stroke, and number of strokes 
per minute; or the last two factors named may be combined in 
the shape of piston speed per minute. The volume of water 
to be raised being given, the cylinder dimensions may be ob- 
tained from lists of standard sizes of pumps, which would usually 
be adhered to on the ground of saving in first cost. With a given 
air pressure and head of water, the diameter of the air cylinder 
obviously depends upon that of the water cylinder. The follow- 
ing relation between the two has been determined by Mr. William 
Cox : f "Area of air cylinder is to area of water cylinder as half the 
head is to the air pressure." By the same writer a ready reference 
table has been constructed, covering the air pressures generally 
used for common, direct-acting pumps: 

* Some suggestive remarks on this subject are made by Frank Richards, 
"Compressed Air," pp. 171-172. 

t Compressed Air Magazine, Feb., 1899, p. 583. (By permission.) 



428 



COMPRESSED AIR PLANT 



Table LI 

Ratios of Diameter of Air Cylinder to Diameter of 
Water Cylinder 









Air Pressure, Pounds. 






Head in Feet. 


















20 


25 


30 


35 


40 


45 


50 


50 


1. 12 


1. 00 


0.91 


0.84 


0.79 


0.74 


0.71 


100 


1.58 


1. 41 


1.29 


1.20 


1. 12 


1.05 


1. 00 


^25 


1-77 


1-58 


1-45 


1-34 


1.25 


1. 18 


1. 12 


150 


1-94 


1-73 


1-58 


1-45 


1-37 


1.29 


1.22 


175 


2.09 


1.87 


1.70 


1.58 


1.48 


1-39 


1.32 


200 


2.24 


2.00 


1.82 


1.69 


1.58 


1-49 


1. 41 


225 


2-37 


2.12 


1-94 


1-79 


1.68 


1.58 


1.50 


250 


2.50 


2.24 


2.05 


1.90 


1-77 


1.67 


1-58 


275 


2.62 


2-35 


2.14 


1.98 


1-85 


1-75 


1.66 


300 


2.74 


2-45 


2.24 


2.07 


1-94 


1.82 


1-73 


325 


2-85 


2-55 


2-33 


2.16 


2.02 


1.90 


1.80 


35" 


2.96 


2.64 


2.42 


2.24 


2.09 


1-97 


1.87 


375 


3.06 


2.74 


2.50 


2.31 


2.16 


2.04 


1-94 


400 


3-16 


2.83 


2.58 


2-39 


2.23 


2. II 


2.00 


425 


3.26 


2.92 


2.66 


2.46 


2.30 


2.17 


2.06 


450 


3-35 


3.00 


2.74 


2-53 


2-37 


2.24 


2.12 


475 


3-44 


^.08 


2.82 


2.60 


2.44 


2-30 


2.18 


500 


3-53 


3-16 


2.89 


2.67 


2.50 


2.36 


2.24 



Ratios for intermediate heads and pressures may be obtained 
by interpolation. 

In this table the unit diameter of water cylinder is taken as one 
inch. Diameters of air cylinders, as calculated, will be in deci- 
mals, and often of odd sizes not occurring in practice. After 
determining the exact diameter, the nearest standard diameter of 
cylinder would be chosen and the air pressure and piston speed 
adjusted accordingly. 

Volume of Air for Pumps Working without Expansion. To 
determine the volume of free air required to operate a direct- 
acting, single-cylinder pump, working without cut-off, the formula 
here given will be found convenient :* 

V = 0.093 ^2 , m which: 

V = volume of free air in cubic feet per minute. . ■ 

h =head in feet under which the pump is to work. 

* Ibid., p. 581. 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 



429 



G = gallons of water to be raised per minute. 
P = receiver gauge pressure of air to be used. 
W2 = volume of free air corresponding to one cubic foot at the 
given pressure, P. '• 
In this formula, which is based on a piston speed of 100 feet 
per minute, fifteen per cent, has been added to the volume of air 
to cover losses. The following table, giving values of W2 and 
0.093 ^2 ^^^ different pressures, may be used in connection with 
the formula : 

Table LII 



Air Pressure P, in Pounds. 


w. 


0.093 W2 


15 


2.02 


0.18786 


20 


2.36 


0.21948 


25 


2.70 


0.25110 


30 


3-04 


0.28272 


35 


3^38 


0-31434 


40 J 


3-72 


0-34596 


45 


4.06 


0-37758 


50 


4.40 


0.40920 


55 


4-74 


0.44082 


60 


5-08 


0.47244 


65 


5-42 


0.50406 


70 


5-76 


0.53568 


75 


6.10 


0.56730 


80 


6-44 


0.59890 


85 


6.78 


0.63054 


90 


7.12 


0.66216 



For example, let it be required to find the volume of free air 
per minute required to raise 200 gals, of water to a height of 
150 ft., the gauge pressure being 30 lbs. From the table, 
0.093 ^2' corresponding to 30 lbs. = 0.2827 ; hence, 

¥ = 0.2827 X ^ = 282.7 cu. ft. free air. 

The horse-power may be calculated from Table LIU, in 
which the mean pressures per stroke (from Table VII), for the 
different terminal pressures, are given in the second column, and 
the calculated horse-powers per cubic foot of free air used, in the 
ihird column: 



430 



compressed air plant 
Table LIII 



Terminal Pressure, Pounds. 


Mean Pressure per Stroke. 


Horse- Power per Cubic Foot 
Free Air, 


20 


14.40 


0.0628 


25 


17.01 


0.0743 


30 


19.40 


. 0847 


35 


21.60 


0.0943 


40 


23.66 


0.1033 


45 


25-59 


0.1117 


50 


27-39 


0.1196 


55 


29.11 


0.1270 


60 


30-75 


0.1340 


65 


32-32 


0.1406 


70 


33-83 


0.1468 


75 


35-27 


0.1527 


80 


36.64 


0.1583 



As the horse-power corresponding to a given terminal press- 
ure does not increase in constant ratio with the initial air press- 
ure, it follows that the higher pressures are not so economical 
for simple pumps as low pressures. Expressed in another way, 
the work of compression decreases with the air pressure, and there- 
fore the useful work done in a pump using air at full pressure is 
greater at low pressures and its efficiency is increased. Thus, in 
the example given above, the horse-power developed in using the 
282.7 cu. ft. of free air, at a pressure of 30 lbs., is: 
282.7X0.0847 = 23.94 h.-p. 

If the air pressure employed were 50 lbs., the cu. ft. of free 
air would be 245.52 and the corresponding h.-p., 29.36, the 
added power cost being 5.42 h.-p. It may be stated that the 
difference in favor of the lower air pressure is offset in part by the 
fact that, at the higher pressure, a pump with a smaller power 
cylinder will do the same work, thus saving in the first cost. 

But the low pressures thus shown to be suitable for simple 
pumps would not serve for machine drills, which must be con- 
sidered first, as they. are in nearly all cases the chief users of com- 
pressed air in mines and quarries. To secure the best results 
from the pumps, a separate, low-pressure compressor would be 
required, a provision which is usually out of the question. Since 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 



431 



it is generally necessary to use high-pressure air, at, say, eighty 
or ninety pounds gauge, the air must either be wire-drawn into 
the pump cylinder or else reduced to the required pressure before 
being delivered to the pump. 

In the first case, the results as to volumes of air used, as given 
in the preceding discussion and tables, must be modified by in- 
troducing a factor of increase, based on the ratio which the press- 
ure to be used in the pump bears to the pressure carried in the air 
main. Edward A. Rix furnishes a table,* part of which is 
abstracted in Table LIV. It shows the volumes of free air 
theoretically required for a unit of 10,000 ft. -gals, of work 
( = 83,000 ft. -lbs. or 2.5 h.-p.), at different air pressures, together 
with the actual air consumption and horse-powers, all referred to a 
standard receiver pressure of 90 lbs. 

Table LIV 



Gauge 
Pressure, 
Pounds. 


Ratio of 
Compres- 
sion, Re- 
ferred to 90 
Pounds. 


Cubic Feet 
of Air Cal- 
culated from 

Cox's 
;; ^Formula. 


Factor of In- 
crease for 
Wire- Draw- 
ing from 90 
Pounds. 


Increased 

Volume, 

Cubic Feet. 


Actual 
Horse- 
Power 
at 90 Pounds, 


Efficiency on 
Basis of 2.5 

Horse-Power 
Theoretical. 


20 


3- 


113 


1.26 


142 


28.6 


9 


25 


2.6 


108 


1.22 


125 


25- 


10 


30 


2-3 


97 


1. 19 


115 


23- 


II 


35 


2.1 


93 


1. 17 


108 


21-5 


II. 6 


40 


1-9 


89 


1. 14 


102 


20.5 


12.2 


45 


1-7 


87 


1. 12 


97 


19.7 


12.7 


50 


1.6 


85 


I. II 


93 


19- 


13-1 


55 


1-5 


82 


1.09 


89 


18.2 


13-7 


60 


1.4 


80 


1.07 


86 


17.4 


14-3 


65 


1-31 


79 


1.06 


84 


16.8 


14.9 


70 


1.24 


78 


1.05 


82 


16.4 


15-3 


75 


1. 17 


77 


1 .04 


80 


16. 


15-6 


80 


I.I 


76 


1.03 


78 


15-6 


16. 


85 


1-05 


75 


1.02 


76 


15.2 


16.4 


90 


I.O 


74 


1.0 


74 


14.8 


16.9 



The factors in column 4 are assumed as about 70 per cent, of 
the ratios of the absolute temperatures due to expansion of the 
air from 90 lbs., to the air pressures in column i. They may be 
taken to apply when the length of air main from the compressor 

* Transactions Technical Society of the Pacific Coast, Aug. 3d, 1900. 



432 COMPRESSED AIR PLANT 

to the pump is moderate, as in carrying the air to a pump situated 
at the bottom of an ordinary shaft. The showing is a poor one, 
but the unfavorable working conditions, as to the type of pump 
and mode of using the air, must be taken into account. 

In the second case, the normal air pressure carried in the 
mine (say, ninety pounds) may be reduced to a suitable pump 
pressure by placing a reducing valve in the air main. The in- 
crease of volume thus produced will be accompanied by a con- 
siderable drop in temperature, so that the full increase is 
not realized. Part of the lost heat will be regained by friction, 
and from external sources if there be any considerable length of 
pipe between the reducing valve and pump; but the efficiency 
will be materially increased if the cold, partly expanded air be 
passed first into an underground receiver and thence to the pump. 
This arrangement has been satisfactorily adopted, for example, in 
the case referred to at bottom of p. 277. An adjustable spring- 
reducing valve is set to furnish any desired pressure below that 
in the main. That is, the volume of air allowed to pass is such as 
to maintain automatically a certain difference in pressure be- 
tween that in the main and the pipe leading to the second receiver. 
The latter serves three purposes: (i) if it be of ample size or of 
the tubular type the air will regain nearly, if not quite, its normal 
temperature; (2) much of the entrained moisture will bedeposited, 
and trouble from freezing avoided; and (3) the receiver, if placed 
near the pump, will minimize the pulsations and equalize the air 
pressure. 

In the particular instance to which reference is here made, 
two underground receivers were installed 300 feet apart, the re- 
ducing valve being put in the main just above the first receiver. 
This arrangement not only caused a very complete deposition 
of the moisture, but the air entirely recovered its normal 
temperature by the time it left the second receiver on its way to 
the pump. The main air pressure was 85 lbs., and at the pump 
about 45 lbs. Indicator diagrams showed 128.5 horse-power de- 
veloped by the compressor and 16.45 horse-power at the pump, 
or an efficiency of 12.5 per cent.; thus agreeing quite closely with 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 433 

the figures in Table LIV. Subsequently, by compounding one 
of the pumps, using 62 lbs. initial pressure in the high-pressure 
cylinder and admitting some live air to the intermediate pipe be- 
tween the cylinders, the efficiency was raised to 25.9 per cent. 
This must be considered a fairly satisfactory performance for a 
pump not specially designed for its work. 

By adopting stage compression or by reheating, or both, the 
total efficiency can of course be increased considerably beyond 
the efficiencies shown in the table. Mr. Rix states, in his article 
previously mentioned, that by actual test of a number of simple 
pumps he has found their work to be approximately 135 
ft. -galls, per cu. ft. of free air. For stage compression the 
efficiency is increased by 15 per cent, (giving, say, 155 ft. -gals.), 
and by reheating the 135 ft. -gals, is increased by the ratio of 
the absolute temperatures under which the pump works, without 
deducting the small cost of reheating. 

Prevention of Freezing of Moisture. Though this subject 
has already been discussed at some length, several additional 
points may be noted in connection with pumping. Some benefit 
may be derived by leading a jet of water from the pump column 
into the air pipe, just before reaching the pump. A very small 
quantity of water will suffice to prevent an excessive drop in the 
temperature of the exhaust. A better way is to tap a one-quarter- 
inch pipe into the column pipe, draw down the end of this pipe to, 
say, one thirty-second of an inch and insert the nozzle so formed 
into the exhaust port. The author has observed the plan of 
carrying a small steam jet close to the exhaust port; but it is 
obvious that this is feasible only when steam is used near-by for 
some other purpose. Moreover, steam so applied is utilized much 
less perfectly than when used in a cylinder jacket. If steam be 
available, a little may be injected into the feed air pipe near the 
pump. An intimate mixture between the steam and air is thus 
produced, and in condensing the latent heat of the steam is given 
up. If water at 212° F. be injected, each pound in cooling down 
to 32° F. will give up 180 thermal units. But with steam at the 
same initial temperature, each pound in condensing gives up 966 



434 COMPRESSED AIR PLANT 

thermal units, in addition to the i8o units imparted in cooling to 
32°. Still another mode of preventing freezing is to warm the 
compressed air by passing it through a coil of pipe, placed in an 
enlarged section of the water column, or else in the pump-suction 
pipe. 

Compressed-Air-driven Compound Pumps. It is a commonly 
held idea that if compressed air be used for operating compound, 
direct-acting pumps, it should be employed like steam, with a 
cut-off in each cylinder. The resulting drop in cylinder tempera- 
ture would be obviously less than that caused in a single cylinder 
by the same ratio of expansion from a given initial pressure. 
But in aiming thus to attain a higher efhciency, by adopting the 
largest possible range of expansion, very low cylinder temperatures 
would still be produced. The loss of heat takes place chiefly 
within the cylinder, instead of in, and just outside of, the exhaust 
port, as is the case with pumps working at full pressure. Fur- 
thermore, though the same total fall of temperature occurs in 
either case, when the air expands within the cylinder the force 
of the exhaust is diminished by the low terminal pressure, and 
the inner portions of the ports are the more liable to be choked 
with ice. 

In order to use the air expansively the necessity for reheating 
in some form is clearly indicated, aside from any question of gain 
in economy. Various plans have been tried of warming the cylin- 
ders by the application of external heat, such as enveloping them 
in a hot-air jacket, surrounding them by water, even heating them 
by the flames of large lamps or torches. But, aside from other 
objections to such devices, air is too poor a conductor of heat to 
render these means at all efficient. 

The mode of applying extraneous heat may be varied in 
several ways, viz: (i) Preheating the compressed air sufficiently 
to permit of a reasonably early cut-off in each cylinder, while still 
avoiding too low an initial temperature in the low-pressure cylin- 
der; (2) in addition to preheating, the air may be reheated be- 
tween the cylinders ; (3) using cold air at full pressure in the high- 
pressure cylinder and expanding into the low-pressure cylinder, 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 435 

with or without reheating; (4) using cold air at full pressure in 
both cylinders, the air being expanded between them, with the 
application of reheating. 

The first two methods are feasible when the compound pump 
is of suitable design and the heating properly applied; but there 
would be an undesirable variation in power and speed, for an 
engine necessarily working under a constant load, if the pump be 
of the usual direct-acting type, without fly-wheel. Moreover, 
under the first plan a high initial temperature would be necessary. 
If the expansion be adiabatic, from an initial pressure of, say, 
eighty pounds to atmospheric pressure and normal temperature, 
the temperature to which the air would have to be preheated is 
given by the expression: 

T' = t(|-')"-^ or, T'-7°°+459°(^°~^^)°''=446°F. 

Although this temperature would be rapidly lowered during 
the stroke, proper lubrication of the cylinder might be interfered 
with. The third method would avoid in part the difficulty of 
variation in power and speed, though there would still be a vari- 
able back-pressure on the high-pressure piston; but the increase 
in volume due to clearance, and on expanding into the passages 
and intermediate pipe to the low-pressure cylinder, would con- 
siderably reduce the temperature of the air, and a large further 
drop would ensue during the work of expansion in the low-press- 
ure cylinder. Such temperature drop may be prevented, or at 
least diminished, by introducing a receiver-reheater between the 
cylinders, with material gain in efficiency. This method has fre- 
quently been adopted, and on the whole is much preferable to the 
two first mentioned. 

The fourth arrangement, however, appears to be the most 
satisfactory. As has been pointed out by E. A. Rix,* in the 
practical application of compressed air to pumps only a small 
part of the total possible work of expansion within the two cylin- 
ders can be realized, even in favorable circumstances. Never- 

* Transactions Association of Engineering Societies, 1900. Mr. Rix also 
proposes the use of three-, and even four-cyhnder pumps. 



43^ COMPRESSED AIR PLANT 

theless, if properly installed and operated, it becomes perfectly 
practicable to drive a compound pump by compressed air. It is 
a much more satisfactory machine than a single-cylinder pump, 
and is capable of working with a fair degree of efficiency. This 
may be accomplished by expanding the air between the cylinders 
only, restoring the consequent loss of pressure by reheating and 
employing full pressure in both cylinders. Thus no drop of 
temperature takes place in the cylinders themselves, and the press- 
cUres, back-pressures, and speed are constant. Each air card is 
practically rectangular in shape. The pressure drop between the 
cylinders may be made small; in fact, it need not be more than is 
sufficient to give the head necessary to cause an active flow of 
air into the intermediate reh eater and thence to the low-pressure 
cylinder. A drop of, say, 20 lbs. for an initial pressure of 70 
to 80 lbs. will usually answer. 

The degree of heat to be imparted by the intermediate re- 
heater, to restore the heat lost by a drop of 20 lbs., would be only 
204° F., for a final temperature of 60° at exhaust. If the pump 
be suitably situated, an ordinary fuel-burning reheater may be 
employed; or, should this be inadmissible, the water from the 
pump-suction or column pipe may be utilized for reheating, as 
already suggested. An example of this arrangement, which has 
often been cited, is to be found in the Gwin Mine, Calaveras Co., 
California.* A Worthington compound pump, having a capacity 
of 200 gals, per minute, was installed on the 600-ft. level of the 
mine. Placed in the suction pipe of the pump is a 300-horse- 
power Wainwright heater, with corrugated copper tubes. The 
water in the pump, at a temperature of 60° to 70° F., passes 
through the heater tubes on its way to the pump-suction valves. 
The air, on being exhausted from the high-pressure cylinder, at a 
pressure of 35 lbs., passes into the heater and through the spaces 
between the tubes. In this way, the temperature of the air is 
raised practically to that of the water and, after expanding again 
in the low-pressure cylinder, is exhausted without freezing. 
Should the sump water be foul, the heater tubes must be cleaned 

* Installed by E. A. Rix. See Engineering and Mining Journal, 1905. 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 437 

from time to time; otherwise the coating of sediment materially 
reduces their conductivity. Still better results would be obtained 
from such an installation by employing a fly-wheel pump with a 
shorter cut-off. The lower temperature could then be met 
by water-jacketing both cylinders, the jackets being supplied with 
water by a small pipe from the pump column. Though the 
quantity of heat thus restored to the expanded air is far smaller 
than that which would be derived from a fuel-burning reh eater, 
this simple device is convenient and satisfactory for under- 
ground service. 

By employing reheating in connection with properly designed 
and operated air-driven compound pumps, efficiencies of 40 to 50 
per cent, may be realized. With 3 -cylinder pumps, furnished 
with intermediate heaters, the efficiencies are still higher, reaching 
even 70 per cent. Reference has already been made to the 
economic advantages of using the Cummings system of high- 
pressure transmission for operating compressed-air pumps. 



CHAPTER XXV 

PUMPING BY THE DIRECT ACTION OF COM- 
PRESSED AIR 

The different modes of raising liquids by the direct pressure 
of air, without the intervention of a piston or other moving part, 
embody no new idea, but it is only in quite recent years that they 
have taken such shape as to render them useful for pumping on a 
large scale. Besides the fundamental considerations of cost and 
efficiency of plant, which affect alike all systems of pumping, an- 
other question becomes of prime importance in connection with 
these methods of applying compressed air, viz: the practicable 
limits of depth or head at which they will work. These limits 
depend on the gauge pressure and mode of using the air. In 
point of efficiency, several forms of plant included under this head 
are distinctly inferior to well-designed steam-driven piston and 
plunger pumps. But when operated under proper conditions 
and with expansive use of the compressed air, recent modifications 
and improvements have brought several of them to a very satis- 
factory degree of efficiency. In first cost they compare favorably 
with pumps of the usual types, and, because of their large capacity 
and low maintenance cost, all possess marked advantages for 
some kinds of service. 

There are two classes of pumps in which the principle in 
question is employed : 

1. Pneumatic-displacement pumps, using compressed air with 
or without expansion. 

2. " Air-lift " pumps, working expansively. 

Pneumatic -Displacement Pumps. These are of several kinds. 
In the type form the compressed air is caused to act directly 
upon the surface of the water contained in a submerged closed 

438 



PUMPING BY THE DIRECT ACTION OF COMPRESSED AIR 439 




Fig. 192. — Merrill Pneumatic Pump. 



440 COMPRESSED AIR PLANT 

chamber or tank, suitable valves being provided for controlling 
the admission of air and water. As the name implies, the water 
is displaced by the air and is discharged from the tank through 
a column pipe. There may be either one or two tanks, the column 
pipe in the latter case being common to both. With one tank, the 
flow of water from the pipe is intermittent; with two, practically 
constant, the pair of tanks then resembling in their relation to each 
other the chambers of the ordinary steam pulsometer pump. 
Aside from the simplicity of construction and absence of moving 
parts subjected to wear, which adapt it for mining, as well as for 
general service, such as pumping from wells and other sources of 
water supply, the pneumatic-displacement pump has a distinct 
advantage for pumping chemical solutions, acids, etc., which 
would corrode the mechanism of a piston pump. It is evident, 
however, that the head or pressure under which the ordinary 
displacement pumps will work is limited absolutely by the air 
pressure employed. 

The double-chamber pump, as built by the Merrill Pneumatic 
Pump Co., will serve to illustrate details of construction and 
operation. Fig. 192 is a diagram of this pump, showing the sub- 
merged chambers, with their connections to the discharge pipe. 
Air from the compressor enters a chest through an automatic 
valve, which opens connection alternately with the two water 
chambers. The air pressure to be employed depends on the 
height of lift. Since the weight of a column of water is 0.434 lb. 
per foot of head, the height to which a given air pressure will raise 
water is equal to the gauge pressure divided by 0.434; thus, air at 

80 
80 lbs. will pump to a height of = 184 ft. In practice, how- 

ever, to cover friction, leakage, absorption of air by the water, 
and to provide the necessary dynamic head for overcoming inertia 
and securing a proper speed of discharge, an additional air press- 
ure is required. In terms of volume, i cu. ft. of water will be 
displaced per cu. ft. of compressed air. One cu. ft. of air at 80 

lbs. = = 6. ^ ^ cu. ft. free air. To this should be added 

15 



PUMPING BY THE DIRECT ACTION OF COMPRESSED AIR 44 1 

for losses, etc., say 20 per cent., making a total of 7.6 cu. ft. free air 

per cu. ft. of water. Taking i gal. of water equal to 0.134 cu. 

ft., the work done per cu. ft. of compressed air, acting against a 

184 

head of 184 ft., will be: - = i8o ft.-ffals. = isos ft. -lbs. 

0.134X7.6 ^ ^ ^ 

In some cases a larger allowance than 20 per cent, should be 
made. The actual work done in compressing i cu. ft. of air to 
80 lbs. gauge, by a single-stage compressor (see Table V) is 0.183 
horse-power, or 6039 ft. -lbs. ; hence, the efficiency of the pump, on 
the basis of allowance for losses assumed above, is nearly 25 
percent., which compares favorably with the efficiencies of single- 
cylinder direct-acting pumps. 

The displacement pump in its usual form works like a simple 
piston pump, in exhausting at each stroke a tankful of air practi- 
cally at gauge pressure. By employing a series of these pumps in 
a shaft, however, and using the air expansively, it is evident 
that, with a given initial pressure, the possible height of lift and 
the total efficiency of the system will greatly exceed that shown 
above.* This can be done by a suitable valve control, by which 
the air is expanded from the lowermost tank to the one next 
above, and so on, for smaller and smaller lifts toward the top of the 
series. When the last tank is discharged, the whole system is 
occupied by expanded air, at a pressure of two or three pounds, 
which is then exhausted into the atmosphere. Air is admitted by 
the valve at intervals into the lowest tank, and the working of the 
system proceeds automatically. At 80 lbs. air pressure, water 
can thus be raised to a height of about 330 ft., instead of 184 ft., 
as in the preceding example, and at an efficiency of about 40 
per cent. 

Another displacement pump is the Latta-Martin, designed 
chiefly for raising large volumes of water under low heads; 
though it may be constructed for any desired air pressure and 
head.f It consists of a pair of submerged cylindrical tanks, 

* This series system of tanks has been proposed by E. A. Rix, Transactions of 
the Technical Society of the Pacific Coast, Aug. 3d, 1900, p. 187. 
t Compressed Air Magazine, Jan., 1907, p. 4332. 



442 COMPRESSED AIR PLANT 

taking water through large disk valves in the bottom. On the 
tops of the tanks is placed the valve mechanism for distributing 
the air alternately into each side. This valve gear comprises a 
main and auxiliary valve, each thrown by a piston valve, similar 
to those of many single-cylinder steam pumps. The movements 
of the valves are caused by the oscillation of a pair of levers, from 
each of which is suspended a bucket filled with water and hanging 
in a housing contained within the main tank. When the pump 
is in operation, the bucket housings are alternately filled and emp- 
tied of water, so that the difference in effective weight of the 
buckets causes them to rise and fall. 

The Harris, or return-air displacement pump, made by the 
Pneumatic Engineering Co., uses the compressed air w^ith some 
degree of expansion. There are two tanks, either submerged or 
within suction distance of the sump, each connected by a pipe 
with the compressor. The water enters by siphon action, the 
inlet, as well as the discharge valves, being placed above the tanks. 
Instead of being exhausted into the atmosphere at each stroke, 
after doing its work, the compressed air is conducted back to the 
intake of the compressor and expands behind its piston. There- 
fore, the system is a closed one, the same air being used over and 
over, in a manner similar to the operation of the Cummings 
return-air plant. The water chambers fill and discharge alter- 
nately, the admission and discharge of the air being governed by 
an automatic switch-valve, connecting the two air pipes close to 
the compressor. 

In starting, after the water in one of the tanks has been ex- 
pelled, the switch reverses and places this tank in connection with 
the compressor intake. Then, while the second tank is being dis- 
charged, the compressed air exhausted from the first returns to the 
compressor and, acting expansively upon the intake side of the 
piston, reduces by so much the power required to drive the com- 
pressor. When the pressure in the first tank has fallen suf- 
ficiently (by being in communication with the compressor intake), 
it will again fill with water. Thus, the compressor transfers the 
same body of air from one tank to the other, additional air to 



PUMPING BY THE DIRECT ACTION OF COMPRESSED AIR 443 

make up for leakage being supplied through an adjustable check 
valve in the intake pipe. This valve is set to open during the 
suction period, at a negative pressure a little greater than the 
pressure required to draw water into the tanks. The switch- 
valve is operated automatically; either by a device acting at the 
intervals required to complete a cycle in both tanks, or by an 
electric make-and-break mechanism, controlled by a pressure 
gauge on the air intake. In the first case it would consist of a 
piston valve, operated by a small air cylinder, compressed air 
being admitted alternately to each side of the piston in the latter 
through an auxiliary valve. The volume of air required for a 
given size of tank may be determined in terms of revolutions of 
the compressor. 

The Harris pump has a high efficiency, say fifty-five to sixty 
per cent., and requires but little attention during its operation. 
It may be adopted for shaft pumping by installing it in several 
units, one above another, according to the total lift. 

The Halsey pneumatic pump is also made by the Pneumatic 
Engineering Co. It has a single, submerged tank, with a simple, 
automatic valve-motion, operated by a float. 

If a displacement pump be required to work in acid water, such 
as frequently occurs in mines containing pyritiferous ore, the 
pressure tanks may be lined with concrete and the other parts 
made of bronze; or the tanks may be replaced by excavations in 
the rock, adjacent to the shaft and lined with concrete or asphalt. 

Air-lift Pump. This, like the displacement pump, is a revival 
of an old principle. Since 1888, in which year Dr. Julius Pohle 
proposed its application for pumping and erected an experimental 
plant, the air-lift has attained considerable prominence. Thus 
far it has been employed chiefly for raising water from deep wells, 
as for water-supply plant, but is applicable to a limited extent 
also for pumping in shafts and for elevating finely divided pulpy 
material mixed with water, such as the sHmes and sands of cya- 
nide and concentration mills. 

The pump consists essentially of. two pipes : a large column 
or delivery pipe and a relatively small air pipe, connected with the 



444 



COMPRESSED AIR PLANT 



compressor receiver. A diagram of the typical form of the ap- 
paratus is shown in Fig. 193. The delivery pipe, open at both 
ends, is submerged in the water to a depth proportionate to, but 
always greater than, the 
height to which the water 
is to be raised. The com- 
pressed-air supply pipe 
passes down to a point 
near the bottom, and ter- 
minates in a nozzle, which, 
directed vertically upward 
by a return bend, is insert- 
ed in the lower open end or 
foot-piece of the delivery 
pipe. (Modifications of 
this arrangement are noted 
hereafter.) 

In some respects the 
operation of the air-lift 
pump is the reverse in 
principle of the method of 
compressing air by the di- 
rect action of falling water. 
As the compressed air 
leaves the small pipe it ex- 
pands and, if the discharge 
pipe is of small diameter, 
tends to form piston-like 
layers, which rise rapidly, 
alternating with masses of 
water. This is readily 
shown by experimenting 
with glass tubes. But if 
the discharge pipe be of 
large diameter, the air 

should be admitted through Fig. 193.— Diagram of Pohle Air-Lift Pump. 




PUMPING BY THE DIRECT ACTION OF COMPRESSED AIR 445 

a series of ports or nozzles, resulting in a dissemination through 
the rising water of small masses of air or bubbles. The water is 
raised chiefly by the buoyancy of the air; or, expressed differently, 
by the aeration of the column of water, which causes a reduction in 
its specific gravity. The action of the pump is due in a small degree 
only to the expansive force and vis viva of the compressed air. It is 
obvious that, before the air is turned on, the water stands at the 
same level inside and outside of the delivery pipe. On entering 
the foot-piece, the air is under a pressure due to the weight of the 
rising column of mixed air and water. As the bubbles of air rise, 
in forcing the water upward, they expand with the decrease in 
head; so that, on reaching the top of the column, the com- 
pression is that due only to the weight or pressure of the small 
quantity of water about to issue from the pipe. Thus, the air 
leaves the pump column at a pressure but little above atmos- 
pheric pressure. The initial air pressure required depends on the 
pressure due to head, measured from the nozzle or air ports 
to the surface of the water. If the pressure be too high loss of 
work ensues at the compressor. Should the delivery pipe be 
too deeply submerged, in proportion to the net height of lift, an 
uneconomically high pressure will be required to force the air into 
the foot-piece; and, with an insufficient submergence a larger 
quantity of air will be necessary to produce the velocity of delivery. 

Referring to Fig. 193, let: 
hi = depth to which the delivery-pipe foot-piece is sunk below the 
normal level of the water, before pumping begins, or when 
the water is at rest; 
/^2 = height at which the water stands when the pump is in opera- 
tion; 
H = height of the column of mixed air and water, measured from 

the air inlet to the point of discharge; 
L = net height of lift=H— /^2. 

The compressed air enters the foot-piece at a pressure, P', 
corresponding to the head, h^; or, 7^2 X pressure per foot of 
hydraulic head =0.434 h^. Assuming that the water rises in piston- 
like masses — as would be the case with a single air nozzle and a 



446 COMPRESSED AIR PLANT 

delivery pipe of small diameter — the sum of the lengths of these 
masses in the column H must be theoretically equal to the outside 
solid column of water, h^. (The weight of the compressed air 
contained in the column may be neglected.) But, to overcome 
the frictional resistance and produce flow, the head h^ must be 
greater. Under ordinary working conditions, the net height of 
lift, L, is found to be from 0.5 h^io say 0.65 h^. Taking the second 

L 

value and transposing: h^=——; and by substituting in the 

L 

expression for the value of P', as above: P' =0.434 —— = 0.67 L. 

0.65 

50 
If, for example, L be 50 ft., P' = 33.5 lbs., and h^ = —- = j'j ft. 

Since the air in the column H is divided into small masses, 
surrounded by water, its expansion during the upward flow may 
be assumed to be isothermal. If P' be its initial pressure, the 

mean pressure for the entire lift = PxNap. log. ( — ),Pand P' 

being absolute pressures. In the above example, taking P as 
15 lbs., P' =33.5 + 15 =48.5 lbs., whence, the mean pressure = 
17.5 lbs. gauge. 

For starting the pump, the air pressure must be suflicient to 
overcome the normal static head, hj, but, when the flow has begun, 
the pressure required falls to that corresponding to h^. Though 
this difference in pressure (hj — /^j) ^^Y be considerable, it is 
readily met by temporarily speeding up the compressor. To mini- 
mize fluctuations between h^ and h^j the top of the well or sump 
should be extended laterally, in order to furnish a large horizon- 
tal area of water, the level of which would be but little affected 
by stoppages or by variations in air pressure and delivery. The 
throttle valve in the air pipe may be regulated by a float on the 
surface of the water. Care should be taken in the design of the 
foot-piece and in properly proportioning the air pressure to the 
submergence and net lift. Otherwise, air may leak back into the 
sump or outside column of water; and, if this becomes aerated, 



PUMPING BY THE DIRECT ACTION OF COMPRESSED AIR 447 

much more power and a larger volume of air will be required to 
keep the pump in operation. In such case the efhciency is greatly 
decreased. 

Since 1889 many experiments by competent engineers have 
been made on the air-lift pump. Among the first were those of 
B. M. Randall and H. C. Behr, on a sixty-foot well, with a stage 
compressor. A summary of these tests is given by E. A. Rix, in 
the Transactions of the Technical Society oj the Pacific Coast, Aug. 
3d, 1900, p. 206. In 1894 a series of tests were made at De Kalb, 
111.,* and in 1893 and again in 1896 on four pumps at Rockford, 
IlLt 

The last-named were carefully carried out and the results 
compared in tabulated form. The heights of lift above water- 
level were 66.5, 90, and 91.5 ft., the air pressure being 76 lbs. 
gauge and the submersion, 225 ft. Both air pressure and depth 
of submersion appear to have been unnecessarily great. With a 
compressor of i24-h.-p., the net work done was 24-h.-p., or 
an efficiency of about 20 per cent. With 600 cu. ft. free air per 
minute, 200 cu. ft. of water were pumped, or 3 air to i water. 
The sizes of piping used were: delivery pipes, 4 in., 5 in., and 6 J 
in., with air pipes from ij to 2 J in. In several of these tests 
the air pipe terminated in a f-inch nozzle. The plan was also 
tried of closing the lower end of the air pipe and discharging the 
air through slot-shaped perforations in the sides near the bottom; 
but the results were inferior to those obtained from the single- 
nozzle opening. Possibly better work would have been done by 
some different arrangement or size of slots; for large pipes and 
volumes of water, at least, the single nozzle has not been found 
satisfactory. 

E. E. Johnson gives a table of the performance of the air-lift 
pump, including consumption of power and theoretical and total 
efficiencies for different height of lift, J from which Table LV 
is abstracted : 

* Engineering News, July 12th, 1894. 
f Ibid., March 4th, 1897. 
J Ibid., April 22d, 1897. 



448 



compressed air plant 
Table LV 







Theoretical 


HORSE-I 


OWER. 


Efficiency of 


Air-Lift. 


Lift In. 


, 














Total Eflficiencv 




P^l 


To Deliver One Cubic Foot 


Thp'^'-Q+i'^il 




from 


Power 






of A 


ir per Mi 


nute. 


J. nt"^'^"^" 






Applied to 




is 














Water 


Del'd. 




1 


i 


1 


.y 


^1 


Si" 


¥4 


rt'35 


3J tj cj 


>/, '^ 




*j c3 


Xi 




a! ?^ 


en's ^ 


■* rt d) 


in 0) 


^'' rt a> 


-^s 


K 


^^ 


S3 


w 


M 


1^ 


6 ex 


^'■^K 


6K 


^^s. 


g 


1 





1 


i 


rt 
^ 


^i 


Hi . 




^1 


•s^i 


fl. 


Uh 


H 


^ 


H 


< 


"^U 


u ^ 


'^ 00 


u 


Woo 


5 


11-54 


.02185 


.02514 


-02572 


.0263 


-87 


848 


-83 


.623 


-497 


lO 


23.09 


-04363 


.05586 


.05992 


.064 


-78 


728 


.684 


-546 


.41 


15 


34-63 


-06545 


.09105 


.0962 


.1015 


.72 


687 


.648 


-515 


•389 


20 


46.20 


.08727 


.12994 


•1391 


.1483 


-675 


.627 


-59 


-47 


-354 


25 


57-75 


.109 


.17191 


.1897 


.2004 


-635 


-575 


-545 


-432 


-327 


30 


69.31 


.13091 


.21678 


-2370 


-2573 


.603 


-548 


.508 


.412 


-305 


35 


80.86 


•1527 


.26445 


-2915 


-3187 


-577 


•52 


.478 


-39 


.287 


40 


92.41 


-17454 


-31375 


-3489 


.3842 


-557 


-502 


■455 


-376 


-273 


45 


103.90 


.1963 


-36368 


-4085 


-4535 


-540 


.482 


■433 


.362 


.260 


50 


115-50 


.21818 


.41848 


-4722 


.5261 


.522 


-464 


■415 


-348 


-249 


55 


127.00 


-24 


.47112 


-5366 


.6023 


-51 


447 


40 


-336 


-24 


60 


138.60 


.26181 


-52855 


.6051 


.6818 


-495 


432 


■384 


-324 


.231 


65 


150.10 


.2836 


.58612 


-6734 


.7608 


-483 


422 


372 


.316 


-223 


70 


161.70 


-30545 


.64812 


.748 


.8483 


.471 


408 


36 


-307 


.2X6 


75 


173-30 


-3273 


.70952 


.823 


.9380 


.462 


398 


35 


-299 


.210 


80 


184.80 


-3491 


-76843 


.898 


I. 0291 


•455 


39 


343 


.292 


.206 


85 


196.30 


■Z1 


-83039 


-976 


1.1231 


-446 


38 


33^ 


-285 


.198 


90 


207.90 


-3927 


.89444 


1-055 


I. 2176 


-439 


373 


324 


.28 


.194 


95 


219.40 


-4145 


.96164 


I-I37 


I. 3148 


-431 


368 


315 


.276 


.189 


100 


230.90 


-43636 


1.0243 


1.247 


1.4171 


.428 


352 


308 


.264 


.185 


110I254.10 


.48 


1. 162 


1-394 


1.626 


-413 


346 


296 


.26 


.177 


120 


277.20 


-5236 


1. 301 


1-571 


1. 841 


.402 


333 


285 


-25 


.171 


130 


300.40 


-5675 


I - 443 


1-755 


2.068 


-394 . 


324 


275 


-243 


-165 



These figures represent the work of well-proportioned plant, as 
to depth of submergence and air pressure. It is shown clearly 
that the efficiency of the air-lift falls off rapidly as the air pressure 
and height of lift increase. The higher efficiencies are naturally 
obtained from stage compression. In general it may be stated 
that, under normal conditions and with small lifts, efficiencies 
of from 30 to 35 per cent, are readily obtainable, and may rise to 
45 or 50 per cent., with proper air pressures and ratios of sub- 
mergence to height of lift. 



PUMPING BY THE DIRECT ACTION OF COMPRESSED AIR 



449 



In 1906 several tests were made at Wandsworth, England, 
on a modified Pohle air-lift, with a delivery pipe of increasing 
diameter toward the top. The total height of the delivery pipe 
was 580 ft., of which 324 ft. were submerged, the net lift thus be- 
ing 256 ft. In this case the distance /Zj — h^ was 69 ft., air pressure 
135 lbs., ratio of volume of free air used to water discharged, 5.8 
and 5.6 .1. The total efficiency was 36 per cent. In view of the 
conditions this is an excellent showing and indicates an advantage 
in using a tapering column pipe. 

The following results of a test made on a 300-ft. well will 
further illustrate this subject:* 

Elevation of discharge above mouth of well 85 ft. 

Depth to water-level during operation of pump 44 " 

Net lift, water-level to point of discharge 129 '' 

Submergence of delivery pipe 248 " 

Air admitted to delivery pipe 5 ft. above inlet end. 

Diameter of delivery pipe 3.5 ins. 

" " air pipe 1.25 " 

Volume of water delivered per minute 82.5 gals. 

" " free air used per minute 81.8 cu. ft. 

Gauge pressure of air 107 lbs. 

Consumption of free air per cu. ft. of water 7.44 cu. ft. 

Horse-power consumed by compressor 12.1"" 

Total efficiency = 22.3 % 

A number of calculated values for air-lift pumps are included 
in Table LVI. 

The question of relative sizes of air and delivery pipes has not 
yet been satisfactorily answered. While there are many varia- 
tions in practice, it is probable that ratios of diameter ranging 
from I : 2 up to I : 2^ or 3 will be found suitable. The absolute 
diameters of the pipes are determined on the basis of frictional 
loss caused by the flow of the air and water. A water velocity of 
250 to 300 feet per minute may be assigned for the delivery pipe. 
The friction losses in air pipes have been discussed in Chapter 
XVI. It should be added that when the water is delivered at a 

* G. C. H. Friedrich, Trans. Ohio Soc. of Mech., Elec, and Steam- Engrs., 1906. 






450 



COMPRESSED AIR PLANT 



distance from the pump, the additional frictional resistance must 
be determined, and the air pressure and submergence corre- 
spondingly increased. Reference may be made in this connection 
to a paper by H. T. Abrams, in Compressed Air Magazine, Aug., 
1906, p. 4135. 

Table LVI 





Volume of Air per 


Submergence, at 




Horse-Power per 


Lift, Feet. 


Cubic Foot 


Sixty per Cent, of 
Total Height of 
Delivery Pipe. 


Air Pressure. 


Gallon Water 




of Water. 




per Minute, 










25 


2 


38 


17 


0.0184 


50 


3 


75 


ZS 


0.0426 


75 


4.5 


"3 


49 


0.0828 


100 


6 


150 


65 


0.1320 


125 


7-5 


188 


82 


0.1910 


150 


9 


225 


98 


0.2544 


175 


10.5 


263 


115 


0-3150 


200 


12 


300 


130 


0.3808 



Among the most complete and valuable recent tests of the air- 
lift pump are those made in 1907 by Messrs. Henderson and Wil- 
son at the two 200-stamp mills of the Angelo and Cason mines, of 
the East Rand Proprietary Mines, Limited, South Africa.* At 
these mills both slimes and sands are raised to the settling tanks 
by air-lift pumps, instead of the usual tailings-pumps and wheels. 
The delivery pipes used in the 19 tests recorded were of two kinds, 
viz: 10- to i6-in. pipes of constant diameter, and several pipes 
increasing in diameter from 12 and 14 ins. at the bottom to 14 
and 16 ins. at the top. These pipes did not taper uniformly, as 
this is impracticable; but, for a length of 35 ft. above the air inlet, 
were lined with one inch of wood, which served incidentally to pro- 
tect the metal from the scouring action of the mixture of sands or 
slimes and water. 

The foot-piece used in the earlier tests was flared out and 
closed at the bottom, the water and pulp being admitted through 
4. large ports, 2 J ft. below the ah* inlet and having a combined 
area of about 200 sq. ins. The air inlet was a single opening, 4 

* The Engineer (London), Jan. loth, 1908, p. 26. 



PUMPING BY THE DIRECT ACTION OF COMPRESSED AIR 45 1 

ins. diameter. For the later tests, the foot-piece was open at 
the bottom and modified by flaring it out to double the diameter 
of the column pipe, so as to increase gradually the velocity of in- 
flow. And, instead of a single air inlet, a ring of twelve holes, one 




Fig. 194. — Foot-Piece for Air-Lift Pump, for Raising Mill Tailings and Slimes. 



inch square, admitted the air ; these holes being cast in an an- 
nular recess a little larger in diameter than the throat of the foot- 
piece. 

The details of the modified foot-piece are given in Fig. 194. 
It is supported on timbers in such a way that the entire bottom 



452 



COMPRESSED AIR PLANT 



is open for the free admission of the material to be pumped. 
The column pipe is of steel tubing, expanded into cast-iron flanges, 
and lined in the lower part with wood, as already stated. This 
design gave materially higher efhciencies than the one first used, 
as set forth in the following table, which, though presenting the 
details of only four of the seventeen tests made, indicates the 
general results obtained. These results show that the air-lift, 
when properly designed for stated conditions, is sufficiently 
efficient to compete successfully with the tailings wheel, in com- 
mon use in the district, and that it is superior to the tailings pump. 

Table LVII 





Test. 


I 


2 


3 


4 




Number and size of delivery 


Two lo-in. 

32-75 
32-5 
1.009 to I 

15 

Original 

TO ins. 


Two lo-in. 

35-75 
29-5 

I. 2 1 to I 

16 

Original 

10 ins. 


One i6-in., 
decreasing 
to 14 ins. 

37-75 

27-5 
1.372 to I 

17 
Modified 

i3i ins. 


One 14-in., 
decreasing 
to 12 ins. 
48-85 
27.09 
1.77 to I 

22 
Modified 


.2 


Submersion in feet 


'« 


Lift in feet 


o 
u 


Ratio of submersion to lift . . . 

Gauge pressure of air, lbs 

Kind of foot-piece 




Throat diameter of foot-piece 


11^ ins. 


S 


Free air, cu. ft. per minute. . . 

" " per cu. ft. of slimes . . 
Cu. ft. of slimes per minute . . 
Throat velocity, cu. ft. per 

second 


2256 
7.27 
310 

4-7 
19-3 

.048 
108.72 


1279 
4.06 

315 

4-8 

17.8 

.050 
64.74 


746.48 
2.74 
290 

4-85 

15-23 

-053 
42.21 


846 

2.64 
320 

7-39 
16.6 




Theoretical horse -power in 
pulp raised 




Horse-power per cu. ft. free 
air compressed 


.064 
54-14 




Air horse-power 








Efficiency, per cent. 


17.7 


27-5 


36-15 


30-55 







In the paper from which the above data are abstracted full 
details of all the tests are given. The conditions were modified 
in the progressive tests, as to the ratio of submersion to lift, 
diameter of delivery pipe, and air pressure. As a basis for cal- 
culating the theoretical horse-power represented by the mixture 
of water and pulp raised, the weight of the slimes was determined 



PUMPING BY THE DIRECT ACTION OF ' COMPRESSED AIR 453 

to be 63.3 lbs., and of the sands, 64.56 lbs., per cu. ft. Thus, 

for the sands, this horse-power was taken to be : 

(Quantity of sands+water)X 64.56 X ft. lift _ Qo^^5y Qyft.ljft. 

33 ^o"^^ 
The term "sands" refers to the mixture of water and ore as 

crushed by the stamps, from which the " slimes " have been 

separated in the milling process. 

LanselPs Air-lift. An interesting modification of the air-lift 
pump, as applied by Mr. George Lansell to pumping water from 
a deep mine shaft in the well-known Bendigo district, of Victoria, 
Australia, may be described here. In the shaft in question water 
has been raised in a series of lifts from a depth of 1,385 feet. Fig. 
195 shows diagrammatically the arrangement of the parts for two 
of the lifts. 

The compressed air is conveyed from the receiver in a pipe, A, 
running down the shaft. The water is conducted from the tank 
or sump through a pipe, D, which first passes down the shaft a 
certain distance, depending upon the height to which the water 
is to be raised, and is then connected with an enlarged section of 
pipe, E, at the foot of the column or delivery pipe, B. Thus, 
the piping for each lift has the form of an inverted siphon, through 
the longer leg of which the water is discharged. At the lowest 
point of the siphon a short branch pipe, C, enters from the air 
main. A, the end of this branch being directed vertically upward 
into the foot-piece, E. Before the compressed air is turned on the 
water stands at the same level in the pipes D and B. The effect of 
this arrangement is similar to that produced by submerging in the 
body of water to be raised the lower part of the delivery pipe, as in 
the Pohle air-lift pump. Check valves are placed, as shown, in 
the pipes D and C, to prevent air or water from passing back into 
the air pipe or into the tank. A throttle valve is provided in the 
pipe C, for regulating the supply of air as required. The relative 
heights of the various parts are not fixed, the dimensions as shown 
on the sketch indicating substantially the proper depth of the 
inverted siphon below the tanks, and the corresponding height of 
lift; thus, from the tank at the 250-ft. level, the pipe D passes 



454 



COMPRESSED AIR PLANT 




Fig. 195- -Diagram of Lansell's Air-Lift Pump for Mine Shafts. 



PUMPING BY THE DIRECT ACTION OF COMPRESSED AIR 455 

down the shaft 140 ft., to the foot of the delivery pipe which dis- 
charges at the surface. A series of lifts may thus be arranged to 
raise the water from any desired depth. The pressure of air is 
the same for all, this pressure being sixty to eighty pounds per 
square inch, or that which is ordinarily furnished for mine 
service. 



CHAPTER XXVI 

COMPRESSED AIR HAULAGE 

For the underground transportation of ore or coal, compressed 
air may be utilized either in locomotives or for driving stationary 
rope-haulage engines. Before taking up the subject in hand, a 
few considerations v^ill be set forth respecting the operation of 
mine locomotives by steam and electricity as v^ell as by com- 
pressed air. Steam locomotives are now much less frequently 
used than formerly for underground haulage, and they can be em- 
ployed only in mines where the trains are conveyed through tun- 
nels or entries directly to the surface, so that stoking may be done 
outside of the mine. Though uneconomical consumers of power, 
steam locomotives are rendered practicable in some collieries 
chiefly by the fact that the fuel is a product of the mines them- 
selves and is therefore chargeable at a low cost. Their principal 
disadvantage lies in the serious vitiation of the mine atmosphere 
caused by the discharge into the workings of the products of 
combustion. Obviously they cannot be employed in gassy or 
fiery mines. 

Electric and compressed-air locomotives divide between them 
a much broader field of operation. Both are applicable to mines 
of all kinds, whether collieries or metalliferous mines; for either 
long or short hauls, from a few hundred feet to several miles; 
they may be used underground in mines worked through shafts, 
where cars cannot be hauled through a tunnel to the surface, but 
must be hoisted on cages, and they do not vitiate the mine at- 
mosphere. For underground haulage in mines containing fire- 
damp, however, electric locomotives must be adopted with cau- 
tion. Although, by the improvements introduced in recent years, 
much has been done to prevent the occurrence of serious sparking, 

456 



COMPRESSED AIR HAULAGE 457 

some risk from this cause still exists; and, furthermore, the 
possibility of strong sparking, accompanied by the momentary 
development of intense heat, from short-circuiting or by reason of 
a ruptured conductor, can hardly be averted. 

Compressed-air locomotives were probably first used in the 
works of the Plymouth Cordage Co., Plymouth, Mass., about the 
year 1873; ^^^ in Great Britain, for mine haulage, in 1878, 
though these early designs were quite different from those now em- 
ployed, and not very successful. Their introduction in the United 
States proceeded very slowly for some years. Perhaps twenty 
compressed-air locomotives were built previous to 1898, but since 
then they have been applied widely for a variety of service.* 
Expressed in general terms, the plant consists of a compressor 
(usually three-stage), receiver, pipe-line, charging stations, with 
the necessary valves and one or more locomotives. The storage 
tank or tanks carried by the locomotive are charged with a suf- 
ficient volume of high-pressure air for a round-trip run of the 
maximum length required, after which the locomotive returns to 
the nearest charging station for a fresh supply of air. 

The special advantages of compressed air, as compared with 
electric haulage for mines, are: First, it may be used in collieries 
with perfect safety, in an atmosphere charged w^ith fire-damp or 
dust, or in dry and heavily timbered workings ; second, since the 
power is stored in the locomotive itself, the system presents the 
maximum degree of flexibility. The locomotives can enter all 
parts of the mine, wherever track is laid, far beyond the limit 
of the supply-pipe line, and are not, like electric locomotives, de- 
pendent upon wiring, which must accompany every foot of ad- 
vance, f For collieries they may be used equally well for the 
haulage of trains on main lines, and for gathering and distrib- 
uting individual cars among the working places; third, the com- 
pressed air costs little or nothing when not in actual use, and its 

* Letter to the author from the H. K. Porter Co., Pittsburg, Pa. 

fit should be noted, however, that storage battery and "cable-reel" electric 
locomotives have been introduced in a few cases, both in Europe and the United 
States. The latter has a very limited range of application and can be used for 
short branch lines only. 



458 COMPRESSED AIR PLANT 

full power or but a fraction of it is available at all times. During 
the unavoidable periods of idleness of the locomotives no power is 
wasted, because, though the compressor may continue in opera- 
tion at a slower speed, it is engaged in storing up power in the 
receiver and pipe-line. Incidentally the exhaust of the locomo- 
tives discharges fresh and cool air into the workings. While this 
is a minor consideration, it improves rather than injures the 
ventilation of the mine. Both electricity and compressed air 
must be looked upon merely as transmitters and distributers of 
power, depending for their production on either steam- or water- 
power as a prime mover. 

At most mines compressed-air haulage is employed only for 
underground transportation, from the stopes or breasts to the foot 
of the hoisting shaft; in other cases, where the mine is worked 
through a tunnel or adit-level, the locomotives haul trains of cars 
direct to the breaker, tipple, or ore-bins, situated on the surface. 
Occasionally, as for example, at the Homestake Mine, Lead, S. D., 
compressed-air locomotives are used for surface transportation of 
ore, from the crusher houses at the shaft mouths to the different 
stamp mills; the object being chiefly to reduce the fire risk for the 
wooden structures, into and near which the haulage tracks pass. 
For the same reasons many plants have been installed in and 
about manufacturing establishments, containing inflammable 
buildings or materials, such as lumber yards and explosives 
factories or magazines. 

Construction and Operation of the Locomotive. For mine ser- 
vice compressed-air locomotives have either one or two cylindri- 
cal storage tanks. These tanks, with the cylinders, piping, and 
other appurtenances, are mounted on a frame provided with 
springs similar to those of a steam locomotive and carried by 
4 or 6 driving wheels. The 6-wheel type is used where a 
heavier locomotive or a lighter rail requires the distribution of the 
load over a greater number of points. Fig. 196 illustrates a recent 
design of a four-wheel, single-tank locomotive, as built by the H. 
K. Porter Co. It is made in several sizes, the details of which are 
given in the first four columns of the following table. 



COMPRESSED AIR HAULAGE 



459 






"'(NO 

? 2 o o 

11 O ^ 



I -2 



t- 2 



1 2 



o B 
o ^ 



O o 

"* 00 



(U 

•S 
U 



Oh 

H 

> 
O 



<u ^ ^^ _ ^ 



Q ^ p p p W 



<D ^ 



1^ 

K o 



J2 '^ 

c ^ 

3 o 

2 ^ 

a, 

^-^ "^ 

O ^3 

c r3 



^ (U 

biD 



C O 



T3 
O 






i 



460 



COMPRESSED AIR PLANT 



In the last two columns of the above table are details of the 
type of locomotive shown in side and rear-end elevation in Fig. 
197. The one illustrated has 4X7-in. cylinders and is used 
on track of 27-in. gauge for hauling mine cars from the under- 
ground loading chutes to the shaft stations. The operator's 
seat is detachable, so that the locomotive can be readily trans- 
ferred as required from one level to another, on a cage whose 
platform is 5 ft. long. These two sizes are suitable for general 




Fig. 196. 



service in metal mines, or for gathering cars from individual 
working places in collieries, to make up trains on main haulage- 
ways. 

A 6-wheel, double-tank locomotive, by the Baldwin Loco- 
motive Works, is shown in Fig. 198. It has the following dimen- 
sions: gauge, 3 ft.; cylinders, 11 ins. X 14 ins.; main tanks, 22 ft. 
7 ins. and 20 ft. i ins. X 34 ins. diameter, carrying a pressure 
of 800 lbs.; auxiliary tank pressure, 140 lbs.; driving wheels, 28 
ins.; wheel-base, total, 6 ft. 6 ins.; total weight, 39,050 lbs., all on 
driving wheels. Another Baldwin locomotive, of the 4-wheel 
type, with 9X 14-in. cylinders, 5-ft. 6-in. wheel-base, and weighing 
24,350 lbs., is shown in Fig. 199. These builders make a number 
of other sizes of mine locomotive, the smallest weighing 8,000 lbs., 



462 



COMPRESSED AIR PLANT 



and having 5JX lo-in. cylinders; track gauge, 36 ins.; tank press- 
ure, 900 lbs., and working pressure 170 lbs. Some of the larger 
sizes are designed for a cylinder pressure of 200 lbs. Compressed 




' — i 


s 


_,,^„^^ 


-f,I..:^.C. UU. -- r« 


SS^Igpfi -. 


^ 


i^ISHkW'^^*^ ^-^^zE 


P^ .V ,,-^r ,4^ j|jjHp5f ^n^ 





Fig. 199. 



air mine locomotives are built also by the American Locomotive 
Company. 

Where there are sharp curves in the track, as is commonly the 
case underground, the wheel-base must be short, say 4 ft. 6 ins. 
to 6 ft., for a 4- wheel engine. The height over all of the loco- 
motive depends somewhat on the conditions existing in the mine 



I 



COMPRESSED AIR HAULAGE 



463 



M rr, ^ 



° 9. 

ro 00 











00 




o 


<N 


ro 












00 


M 


-t-> 




rO 




00 


rH|(M 














00 


00 






M 


ro 












Tl- 


0^ 








'"' 


vO 


Ml-* 








ro 








O rO 



o 
o 
o •^ 



X 



W 



.8 



,L 5 5 



o o 

10 o 

H 00 



o o 

fO o 

H 00 






u 
19 



OJ -H -k-" 






:3 3 3 

CO t/3 t/3 

p p 13 






c 
o 



■ w 






Oh 
< 



o "d 






> 


> 


;h 




;3 


s 


u 













<u 


(U 


Oh 


d. 






«5 


03 



5 Oh 



a, 
< 



r/l 



.S CI. 

cr aj 



;3 :^ 
!^ .1:3 



464 



COMPRESSED AIR PLANT 



as to thickness of vein, head-room of the haulageways, etc., and is 
rarely more than 5 or 6 ft. — frequently less. The length varies 
greatly, mainly according to the tank capacity required, and the 
curvature of the gangways. It is usually from 10 to 15 ft. for 
the smaller sizes, up to 20 or 24 ft. for the larger, the widths 
ranging from 3 J to 6 ft. 

Table LIX contains the principal data of seven sizes of 
large six-wheel, double-tank locomotives, built by the H. K. Porter 




Fig. 



Co. These comprise the heaviest locomotives designed for under- 
ground mine service. 

Additional details of the construction of compressed-air mine 
locomotives are exhibited in Fig. 200, containing a general plan and 
side elevation of a Baldwin locomotive, with 9Xi4-in. cylinders. 
Fig. 201 shows a half front-end elevation of the same, with half 



p 

-/ 

-V 



r^-i 



i^i}iiim4fLU-nn 



lilW r i -, 1 II 



// 



/ /r-\ lil ' ' I iih 



\-H-^] 



ill I 1 1 1| I 
I I 



itT^i' TTTrrTTTi 

-/^ — rAr 1^ — ^ 



Ivl-4 



\ 

'I \ 



— V- 



.j-V-t— i-^ 1- 



w 



\ \ 



i[?i rif^ 



Itl 



\ \ "[ij I ) 1 u Lij u J I I II ; 



III' 




Fig. 200. — Baldwin Compress' 



7F^ P" 



4- 



''''-^^--\^A W 



-13 >i — >i 






I [ \ 



u 



vf 



' '^^ ^TTl 



^1 V\ \ 



n 






I I L .■ I.' /? / 



/ / 










Itive. (Plan and Elevation.) 



(.noijfl 



COMPRESSED AIR HAULAGE 



465 



section through frame and left-hand storage tank; and Fig. 202, 
a half rear-end elevation, with section of left-hand tank. 

The tanks have dished or approximately hemispherical ends, 
and are built of extra heavy steel boiler plate; the shells being 
f to I in. thick, with i to ij-in. heads. Ring seams are double 
riveted with lap joints; longitudinal seams being butt joints, with 
inside and outside welt strips. As the tanks are generally built 
to carry working pressures of 700 to 800 lbs. per sq. in., the longi- 




FiG. 202. 



tudinal seams have 6 to 8 rows of rivets, to make a joint of not less 
than 75 per cent, of the strength of the plate. It is customary to 
test the tanks to 800 or 1,000 lbs., the factor of safety with plate 
of the usual quality b^ing, say, 3 J. This is considered sufficient, 
as there are no strains produced by expansion and contraction, as 
in a boiler. When extremely high pressures are required, tanks of 
large diameter cannot safely be employed, and are replaced by a 
set of heavy seamless steel tubes, 8 top ins. diameter — for example, 
the Mannesmann tubes. Tubes of this kind, 9 ins. diameter by 



466 COMPRESSED AIR PLANT 

^ in. thick, will carry working pressures of 2,000 to 2,500 lbs. per 
sq. in. A number of them are laid together, bound by belts or 
straps and then enclosed in a light sheet-iron shell, to protect them 
from wet and rust. But these high pressures are unnecessary for 
ordinary systems of mine haulage. 

From the main tanks the air passes into a small auxiliary or 
distributing reservoir and thence to the cylinders. This auxiliary 
tank is merely a section of wrought-iron pipe from 4 to 9 ins. 
diameter and 6 to 15 ft. long, with closed ends and laid alongside 
the main tank. By means of an automatic reducing valve, the 
pressure in the small reservoir is adjusted to the requirements of 
the engine. As used on the locomotives of the H. K. Porter Co., 
the reducing valve consists of a double-seated balanced valve, 
operated by a small piston. The air pressure in the auxiliary 
reservoir acts on one side of this piston and tends to close the 
valve. This action is opposed by a powerful external spring, 
which is adjusted to keep the valve open until the normal working 
pressure is reached in the auxiliary reservoir. Then the valve is 
closed by the air pressure, against the resistance of the spring. To 
provide for the case when the locomotive is using no air (as on a 
down grade or when at rest), a single-seated supplementary valve 
is placed in the pipe between the reducing valve and the loco- 
motive storage tanks. This valve is controlled by the throttle 
lever; being open when the throttle is open, otherwise closed by 
the air pressure. By thus using two valves leakage from main 
tanks to auxiliary reservoir is avoided and a close regulation 
secured. 

The cylinder pressure adopted ranges generally from 125 to 
150 lbs., according to the size of cylinder and power required, thus 
being about one-quarter of the pressure in the main tank. From 
the small tank the air passes to the cylinders through a balanced 
throttle valve. This arrangement permits the maintenance of a 
constant working pressure, suited to the needs of the locomotive, 
prevents the waste of aii likely to ensue if air at full tank pressure 
were admitted to the cylinders, and makes the locomotive more 
manageable. The cylinders, moreover, need not be made so 



COMPRESSED AIR HAULAGE 467 

heavy as would be required for a high pressure. In starting a 
heavy load excessive slipping of the drivers is avoided, and with 
light loads the reducing valve may readily and quickly be regu- 
lated to produce any desired reduction of pressure. In the opera- 
tion of the locomotive toward the end of the haul, when the press- 
ure in the main tanks falls to that in the auxiliary tank, the 
cylinders take their air directly from the former, and the loco- 
motive w^ill continue to run as long as the pressure remains 
sufficient. Sometimes, for long hauls, and when the cross- 
sectional dimensions or sharp curves, or both, of the haulage- 
way do not permit the use of tanks of great length or large 
diameter, a tender carrying a supplementary tank is employed. 

For small-scale work, the air is sometimes admitted to the 
cylinders throughout nearly full stroke, and consequently, as the 
exhaust is at high pressure, the efficiency is lower than it should 
be. This practice is doubtless due to the tendency to use as 
small a motor as possible for the service required, on account of 
the limited head room and narrow, crooked gangways so common 
in mines. Better results are obtained by using a cut-off and in- 
creasing the size of the locomotive and the weight on the drivers. 
This is almost always done with large locomotives. Ample re- 
serve power is available when necessary, since full tank pressure 
can be temporarily admitted to the valve-chests in starting a 
heavy load, or in hauling on steep grades and around sharp curves. 
In using the air expansively, as can be done with properly pro- 
portioned cylinders, there should be no trouble from freezing of 
the moisture. Although the expansion will produce a low cylin- 
der temperature, yet, as the initial working pressure is so much 
higher than is employed for pumps or other compressed-air 
machinery, the expanded air becomes relatively dry, and the force 
of the exhaust is still sufficient to keep the ports clear of accumu- 
lated ice. To this end the ports should be large, straight, and 
short, though ports of ordinary proportions are quite common. If 
high-pressure air were used in the engines, both cylinders and 
pistons would have to be made excessively heavy, and any reason- 
able degree of expansion would produce a degree of cold difficult 



468 COMPRESSED AIR PLANT 

to deal with. The cylinders should not be lagged with non-con- 
ducting covering, as is so necessary for steam cylinders, to min- 
imize condensation. By exposing their surface to the warm 
air of the mine, some heat is absorbed. Usually the exterior 
surface of the cylinders is cast with deep corrugations, in order to 
present the largest possible superficial area to the warm sur- 
rounding air. The cylinders are provided with slide valves; 
piston valves, like those used in steam locomotives, would leak 
more because of the dryness of the air. 

On account of the cold produced by the reduction of pressure 
from the main tanks to the auxiliary reservoir, and to increase 
efficiency of operation, reheating is found to be advantageous, 
though not essential. It may be accomplished conveniently by 
applying heat to the auxiliary reservoir. If steam be available in 
the mine, a quantity of steam and hot water may be injected into 
this reservoir each time the locomotive is charged. Or, in mines 
where there is no danger from fire-damp, a small reheating ap- 
paratus for burning oil or coke may be carried on the locomotive. 
It is always desirable to warm the reducing valve from the main 
tank, as this is subjected to intense cold. In any case, when the 
air is reheated a quantity of water should be kept in the small 
tank. An incidental advantage of this arrangement is that the 
moisture from the hot water, which passes with the air into the 
cylinders, assists in lubricating the valves and pistons.* 

Pipe-line and Charging Stations. The capacity of the com- 
pressed-air system naturally depends on the length of haul and 
size of locomotives, as influenced by the daily output, weight of 
trains, and gradients of the haulage lines. For short hauls, the 
pipe-line is sometimes omitted altogether, the locomotive return- 
ing each time to the compressor receiver to be recharged. In 
general practice, however, a pipe-line is carried underground, and 
at one or more points charging stations are established. The lo- 
cation and distance apart of these stations is determined by the 
haulage distances and the storage capacity of the locomotive 



* E. P. Lord, Paper Read before the Anthracite Coal Operators' Association, 
N. Y., Oct. 13th, 1897. 



» 



COMPRESSED AIR HAULAGE 469 

tanks. It is evident that the last or innermost charging station, 
farthest from the compressor, must be at a point from which the 
locomotive can reach the end of its trip and return for a fresh 
supply of compressed air. For very long hauls, heavy trafhc, 
or adverse gradients, a charging station may be required at each 
end of the line. 

It is unnecessary to provide receivers inside the mine, though 
this may be done advantageously if the diameter of the supply 
pipe is small. The pipe-line itself is intended to act as a storage 
reservoir, and should be of a diameter which, in proportion to its 
length, will furnish a cubic capacity sufficient to charge the 
locomotive tanks quickly and without serious drop in pressure. 
In other words, when the locomotive is connected with the pipe- 
line, and the charging valve opened, the pressure in the locomotive 
tank and in the pipe, on equalizing (as it must), should not fall 
much below the stated pressure which the locomotive is designed 
to carry. It is, therefore, desirable that the volume of storage, 
represented by the main — or main and receiver — should be at 
least three times the tank capacity of the locomotive. To de- 
termine the necessary storage capacity of pipe-line, or combined 
receiver and pipe-line, several variables must be harmonized, as 
follows:* 

V = storage volume required, in cu. ft. 

V = volume of locomotive tanks, in cu. ft. 
P = pipe-line pressure, in lbs. per sq. in. 

p = desired pressure in locomotive tanks, in lbs. per sq. in. 
_/>'= residual pressure in locomotive tanks, just before charging, 
in lbs. per sq. in. 

Then: V (T-p)=v (p-p'), or V= ^^^ 

For example, let P=9oo lbs., p = 7S^ lbs., ^ = 125 lbs., and 
f = 100 cu. ft., from which: 

__ 100 (750—125) ^ ^ r. 

V = 1^^ ^' =416.6 CU. ft. 

900-750 

By transposition, the same formula may be used for finding 

* H. K. Porter Co., " Handbook of Compressed-Air Haulage," 1907. 



470 COMPRESSED AIR PLANT 

the pipe-line pressure required to produce a given pressure in 
the locomotive tanks. When several locomotives are served by 
the same pipe-line and compressor it is rarely, if ever, necessary 
to design the system for charging more than one at a time. If 
the volumetric capacity of the pipe-line be ample, the relatively 
small drop in gauge pressure on charging is soon recovered by 
the compressor, which, except in plants operating a single locomo- 
tive, is kept in nearly constant operation. In case additional 
locomotives are required after the original installation of the 
system, the same pipe-line may still serve, provided the com- 
pressor be of sufficient size to charge it to full pressure at shorter 
intervals. 

The piping, which generally varies in diameter between 3 
and 5 ins. — sometimes 6 ins. — should be of the best material, lap- 
welded, and with sleeve joints made with the utmost care to pre- 
vent leakage. To stop leaks, the sleeves should have annular 
grooves at each end into which soft metal calking is driven if 
required. It is advisable not to bury the pipe alongside the 
track, but to ca-rry it entirely uncovered along one side of the tun- 
nel or gangway, either on the floor or on brackets, so that leaks 
will at once attract attention and be stopped. While an occa- 
sional bend in the pipe-line is advantageous in permitting free 
expansion and contraction, they should not be too numerous, as 
they involve more joints and therefore a greater possibility of 
leakage. 

Charging Apparatus. A common form of apparatus for charg- 
ing the locomotives, as shown in Fig. 203, consists of a vertical 
right-angled connection inserted in the air main by means of a 
heavy tee. This connection has an arm projecting from the main 
a sufficient distance for conveniently coupling to the charging 
pipe of the locomotive. It comprises two parts: a vertical, 
rigid branch, containing a strong, accurately fitted i J- inch gate- 
valve, and a short horizontal pipe, attached to the valve by a 
union and a ball-and-socket or flexible joint, for coupling to the 
locomotive charging pipe. Thus, the locomotive need not be 
stopped at a precise point for charging, but has a foot or two lee- 



COMPRESSED AIR HAULAGE 



471 



way on its track. When not in use, the flexible connection is 
s\\Ting back, out of the way. In the locomotive connection there 
are usually two ball-and-socket joints, together with a check- valve 
close to the tank. 

After coupling on the locomotive, the gate-valve is opened, 
whereupon the air pressure immediately forces together the parts 
of the ball-and-socket joints and makes a perfectly tight connec- 




d 
'to 

U 



CO 

o 
w 

6 



472 



COMPRESSED AIR PLANT 



tion. As soon as equilibrium is established between the pressures 
in the main and the locomotive tank the gate-valve is closed. To 
break the coupling, the compressed air remaining in the connect- 
ing pipe, between the gate-valve and the locomotive check-valve, 
must first be released. This is done by opening a small "bleeder 
valve," placed just above the gate-valve, as shown in the cut. 
The joints then become loose and are readily manipulated. The 
actual time occupied in charging is very short (usually about 
three-quarters of a minute), owing to the high pressure in the 
main and the relatively large diameter (ij in.) of the charging 
pipe; but, including stopping the locomotive and making the 
connection, i J to 2 J mins. may be allowed. Frequently, charging 
may be done during the necessary delays in shifting cars and 
making up trains. 

Calculation of Motive Power. To determine the motive 
power required for a given output, several factors must be known, 
vi/.: the tractive resistance per ton of the loaded cars on a level, 
the resistances due to gradients and curves, the weight of empty 
and of loaded cars, and the number of cars to be hauled in each 
train. The values of these factors are known approximately or 
are readily ascertained, with the exception of the resistances due 
to curvature of track and character of roadbed. The former has 
been determined experimentally for ordinary surface railways^ 
but underground mine track is apt to be roughly laid, with curves 
of varying and irregular radius, and the elevation of the outer rail 
improperly adjusted. With sufiicient weight on the drivers^ 
however, sticking on a curve may be avoided, in the case of com- 
pressed-air haulage, by temporarily admitting to the cylinders a 
little air at full tank pressure, as already noted. In this respect 
compressed-air locomotives possess a material advantage over 
those driven by steam, in which the working pressure is limited 
and practically constant. 

The average tractive force required per ton depends not only 
on the physical condition of the track and roadbed, but on the 
character and state of repair of the running gear of the cars. On 
level mine track the coefficient of rolling friction should usually 



COMPRESSED AIR HAULAGE 473 

be taken at from thirty to forty pounds per ton, though it may 
be considerably higher on poorly laid or light track, or at the in- 
stant of starting the load. With mine track in exceptionally good 
condition, the coefficient may be as low as twenty pounds per ton. 
The grade resistance is twenty pounds per short ton, for each one 
per cent, of grade. Not infrequently, the distribution of grades on 
the haulage lines is such that the maximum load is not the resist- 
ance of the loaded trains, which are usually hauled on slight down 
grades, but that of the return trains of empty cars on the adverse 
gradients. To obtain the most economical results, gradients 
should be not over Jto fof i per cent, in favor of the loaded trains. 
With mine track and rolling stock of ordinary character, and a 
grade of 5 to 6 ins. per 100 ft., the coefficient of rolling friction is 
nearly the same for a loaded train hauled down as for an empty 
train of the same number of cars hauled up the grade. Heavier 
and even adverse grades often become necessary — sometimes as 
steep as 2 J per cent, to 3 per cent, or more, but they should be 
avoided as far as possible, because the maximum tractive force of 
the locomotive falls off rapidly. On a 2 J-per-cent. adverse grade 
the locomotive can haul only about 4 times its own weight, even 
if the track be not slippery. Grades should be reduced on curves. 
Colliery cars, carrying 2 J to 3 J tons, will weigh from 1,800 to 
2,300 lbs., while those used in metalliferous mines, where me- 
chanical haulage is employed, vary between, say, 1,000 and 2,000 
lbs. Many cars of the last-named weight are in use, for example, 
in the iron mines of the Northwest. Finally, having ascertained 
as near as possible the values of the different factors, the proper 
allowance of reserve power, in terms of volume and pressure of 
air, to cover indeterminate additional resistances due to imper- 
fections of track and rolling stock, is a matter of judgment and 
experience. 

With a given air pressure, the capacity required for the loco- 
motive storage tanks depends primarily on the length of round 
trip to be made with a single charge of air. When this distance is, 
say, I to I J miles, the tank capacity generally varies between 50 
and 150 cu. ft., according to the load; which, in turn, together 



474 COMPRESSED AIR PLANT 

with the track and grade resistances, governs the dimensions of 
the cylinders. Cylinders of 5 ins. X 10 ins. up to 9 ins. X 14 ins. 
are commonly used for mine service, the larger sizes being 
adopted for heavy work in collieries. Still more powerful locomo- 
tives are used for some kinds of surface work. In several installa- 
tions, as at mines of the Philadelphia & Reading Coal & Iron Co., 
the compressed-air locomotives have been designed with com- 
pound cylinders. For long runs, of over one and one-half miles, it 
is often desirable to increase the air pressure, rather than build 
tanks of very large size. Another plan is to provide a tender, 
which carries one or more auxiliary tanks, connected with those 
on the locomotive. Very long runs can be made by this means. 

Having determined the total work in foot-pounds to be done 
with a single charge of air, on a run of the maximum length, 
specifications may be obtained from the builders for a loco- 
motive of suitable weight, gauge, wheel-base, tank capacity, and 
cylinder dimensions. 

Compressors for Charging Pneumatic Locomotives. For 
compressing the air to the high tension required by pneumatic 
locomotives, the work must be done in at least 3 stages; 
4-stage compressors are sometimes employed for pressures ex- 
ceeding 900 or 1,000 lbs. Efficient intercoolers are of course 
placed between the successive cylinders and an aftercooler is 
desirable. Fig. 204 shows the standard type of 3-stage locomo- 
tive charger built by the Norwalk Iron Works Co., for pressures 
up to 1,000 or 1,200 lbs. The air passes from the low-pressure 
cylinder to the lower of the two intercoolers and, thence to the 
intermediate cylinder. From the latter the air is delivered 
through the vertical pipe to the upper intercooler, whence it 
passes through the inclined pipe to the high-pressure cylinder. 
From this cylinder the compressed air is delivered to the receiver 
through the connection indicated under the outer end of the 
cylinder. Other compressors by the same builders are designed 
for pressures up to 3,000 and 4,000 lbs. 

The air end of a three-stage, tandem locomotive charger, 
built by the Ingersoll-Rand Co., is shown in longitudinal section 



COMPRESSED AIR HAULAGE 



475 



in Fig. 205. The high-pressure intercooler is placed in the lower 
right-hand corner of the cut. Figs. 206 and 207 illustrate respec- 
tively the low- and high-pressure air ends of a duplex, four-stage 
compressor. In Fig. 206 are the intake and first intermediate 
cylinders, and in Fig. 207 the second intermediate and high-press- 
ure cylinders. A perspective view of a large compressor of this 
type is show^n by Fig. 208. 

It will be seen in the sections that the pistons of the high-press- 
ure cylinders are solid rams or plungers, provided with a series 





i u- ' ,„^ 


'#"1 


i^^'~ 


1 


c 


M — y^SfSS^SBSSSSSSSSBSBsmgm 




}: W^M^-m K -^ ,... mm-Mtif.: 




m^L 


ti^mClM. 






%%r^^^P-- 


;--^-'Pi?|t ■'::■;.,:- 


PWI^^mirf^ 




m^gj^^Sst * .ri^ -s .Z^.^^^ 




^*'*" -mi'm '\ 


W^^^^Bm,'i 



Fig. 204. — Norwalk Locomotive-Charging Compressor, 

of packing rings. These, with the high-pressure valves, must be 
made with special care, to prevent the serious effects of leakage 
of high-pressure air. Even a small percentage of leakage will 
greatly reduce the volumetric capacity and efficiency. Loco- 
motive chargers are also built by the Sullivan Machinery Co. 
and others. 

When the mine is already provided with an ordinary low- 
pressure air plant, for machine drills and other service, and which 
has some surplus capacity, a two-stage charging compressor may 
be installed, to take air from the low-pressure system and bring it 



476 



COMPRESSED AIR PLANT 




COMPRESSED AIR HAULAGE 



477 




Figs. 206 and 207.— IngersoU-Rand Four-Stage Locomotive Charger. 



478 



COMPRESSED AIR PLANT 



up to the tension required for the locomotives. By this arrange- 
ment some reduction in the cost of the plant may be effected. 
Care must be exercised, however, in making such a combination. 
If the quantity of air produced by the low-pressure system should 
at times be insufficient to furnish the necessary excess, at ordinary 
gauge pressure, for the locomotive-charging compressor, the latter 
might be compelled to compress from too low an initial pressure. 




Fig. 2oi 



This would cause excessive development of heat and, aside from 
the difficulty of maintaining proper lubrication, might possibly 
raise the cylinder temperature to the flashing-point of the 
oil, thus causing an explosion. This matter has been discussed 
in Chapter XIV. Generally, it is preferable to install an inde- 
pendent locomotive charger. With such a compressor, the final 
temperature can be kept down to a moderate degree, say, 200° 
to 300° F., provided the plant is not too small for its work. The 
compressor should be run at a moderate speed, and as the de- 
mand upon it is usually somewhat irregular, causing frequent re- 
ductions of speed, and even occasional stoppages, the air cylinders 
are prevented from becoming over-heated. 



COMPRESSED AIR HAULAGE 479 

The capacity of the charging compressor depends on the pipe- 
line pressure to be maintained, the number of locomotives to be 
operated, the cubic contents of the locomotive tanks, the pressure 
carried by the system, and the relation between the charging 
periods. 

Let C= compressor capacity required, in cubic feet of free air 
per minute. 
L= locomotive-tank capacity, in cubic feet of free air per 

minute. 
N = number of charges required in any given time, T. 

. ^ NL 

Hence the equation: C = -^=^ 

For example, if N = 3, L = 5,200 (corresponding to 100 cu. 

ft. of air at 750 lbs. gauge pressure), and T = 60 minutes : 

„ 2X^,200 , , . . . , 

C = ^^ — ^ =260 cu. ft. free air per mmute. 

60 

When the locomotives are charged — as they usually can be — 
at approximately equal intervals of time throughout the day, a 
single application of the above formula will be sufficient. Other- 
wise, calculations are required to determine the maximum and 
minimum rates of consumption of air. It is hardly necessary to 
add that, when the plant is installed at an altitude above sea- 
level, allowances must be made for decreased output, as ex- 
plained in Chapter XIII. 

Examples of Compressed-Air Haulage Plants. In further 
illustration of this subject, some of the details of a few successful 
installations may here be given. 

I. At the Buck Mountain Colliery, Penn., are two 8-ton H. 
K. Porter locomotives, each with 2 tanks, respectively, 15 and 17 
ft. long, having a combined capacity of 130 cu. ft. of air at 550 
lbs. pressure.* The cylinders are 7 ins. X 14 ins.; wheel-base, 5 
ft. 3 ins. ; gauge of track, 42 ins. ; height, 5 ft. 2 ins. ; length over 
all, 19 ft. A round trip of 5,100 ft. is made in 30 to 40 minutes, 
or 2,500 ft. in 12 to 15 minutes, with trains of 12 cars, on grades of 
J to 4I per cent., averaging | of i per cent, in favor of the load. 

* Mines and Minerals, July, 1898, p. 538. 



480 COMPRESSED AIR PLANT 

One locomotive delivers 150 cars per 10 hours, doing the work 
formerly done by 15 mules. Weight of cars, 3,400 lbs. empty, and 
10,400 lbs. loaded. A 3 -stage Norwalk compressor supplies 
375 cu. ft. free air per minute, at 700 lbs. gauge. Pipe-line, 4 ins. 
diameter and 9,600 ft. long, with a storage capacity of 800 cu. ft. 

Average cost per ton-mile: 1.875 cents for the gross weight 
hauled, or 3.77 cents for net weight of coal. The cost for mule 
haulage under the same conditions was formerly 3.94 and 7.92 
cents, respectively. 

The cost of this plant was as follows: 



Two locomotives ^5>505 ■ 

Air line: 9,647 ft. 4 in. pipe $2,894. 

Six charging stations 360. 

Fittings and valves 382 , 

Labor cost for erection 998. 

4.634. 



Compressor $2,880 

Sundries and erection 246 

Compressor house 256 

Steam line to compressor 152 



3.534. 



Total cost $13,673. 

2. Empire Mine, Grass Valley, Cal. Several small com- 
pressed-air locomotives, built by Edward A. Rix, are employed in 
the deep levels of the mine, for hauling trains of 5 cars, each carry- 
ing I ton. The maximum distance covered by a round trip is 
about 5,000 ft. Locomotive storage tank measures 36 ins. 
diameter X 48 ins. long, carrying a pressure of 500 lbs. The di- 
mensions over all are only 5 ft. long X 30 ins. wide X 52 ins. high, 
the gauge of track being 18 ins. One of these locomotives (Fig. 
209) is operated by a pair of vertical engines, a chain and sprocket 
drive connecting the crank-shaft with the rear axle. There are 
2 tandem tanks, one of them being carried on a tender. A re- 
heater, provided with a Primus kerosene burner, reheats the air 
after its pressure has been reduced in the auxiliary reservoir. 
Mr. Rix has recently built 3 similar locomotives, but with a single, 
larger tank, for a 3-mile tunnel, near San Francisco. They carry 



482 COMPRESSED AIR PLANT 

1,000 lbs. tank pressure, the working pressure being 100 lbs.; 
each locomotive making a 2 -mile round trip, at 6 to 7 miles per 
hour.* 

3. The Peerless Colliery, Vivian, West Va., operated for 
years several H. K. Porter locomotives, with 5Xio-in. cylinders 
and weighing 10,000 lbs. Over all dimensions : 10 ft. 5 J in. long 
X 5 ft. 8 ins. wide X 4 ft. 5 ins. high. Four driving wheels, 23 ins. 
diameter ; gauge, 44 ins. Capacity of main storage tank, 47 cu. ft. ; 
pressure, 535 lbs.; charging time, 20 seconds; working pressure, 
125 lbs. Pipe-line, 3 ins. diameter, with a total capacity of 242 
cu. ft. Line pressure, 735 lbs. Trains consist of 6 cars, each 
weighing loaded 8,500 lbs. Grades range from level to 2 J per 
cent., generally in favor of the load. Curves from rooms to haul- 
ageways, 23 ft. radius, though locomotives are designed to work on 
curves as sharp as 15-ft. radius. Length of maximum round trip, 
9,000 ft. ; maximum speed 10 to 12 miles per hour. Cost of each 
locomotive, $1,800. 

4. The following data, concerning one of the plants of the 
Philadelphia & Reading Coal & Iron Co., and compiled by Mr. 
G. Clemens, a division engineer of the Company, are published 
in the catalogue of the Baldwin Locomotive Works : 

a. Shaft level — i locomotive. 

Round trip, 5,200ft. ; grades y^toy^ of i per cent., all in favor 
of load; charging station at each end of run; gauge of track, 44 
ins.; 40-lb. rails; weight of cars — empty, 3,300 lbs., loaded, 8,800 
lbs.; from 15 to 38 cars per trip; total output, 600 cars per 10 
hours. Round-trip time, 12 miri.; charging time, i min. A 
round trip and a half can be made with one charging. 

h. Slope level — i locomotive. 

Length of haul, 3,200 ft., of which 700 ft. is on a slope whose 
grade ranges from /\.-^-^ to 5^ per cent. Grade of main gangway. 
To" t^ tV^^ ^ P^^ cent., in favor of load. Trains of 10 cars are 
hauled on main gangway, and 4 cars on the slope; weights of cars 
same as above. 

Locomotive-tank pressure at start, 600 lbs. ; at end of trip, 

* Compressed Air Magazine, Feb., 1908, p. 4747. 



COMPRESSED AIR HAULAGE 483 

200 lbs. Average working pressure, i8o lbs. The cost of the 
plant was as follows : 

)ne Norwalk 3-stage compressor, erected $5, 180. 74 

ipe-line, 4,200 ft., 5 in., including 3 charging stations 2,951.06 

Two Baldwin compressed-air locomotives and fittings 4,904.33 

Alterations in gangways to adapt them for locomotive haulage 665 . 1 7 

Total cost $13,701 .30 

Daily operating cost, for 180 days in the year $14 . 69 

Fixed charges, depreciation, repairs, etc., figured at 10 per cent., together 

with cost of steam power 9 . 00 

Total running expenses per day $23 . 69 

Cost per car, at 660 cars per day 3.6 cents 

Previous cost of mule haulage per car 5.1 " 

Saving per year, about $1,800.00 

5. At the Wilson Colliery, of the D. & H. Coal Co., a large 
locomotive was installed by the Dickson Manufacturing Co., hav- 
ing six 26-in. drivers; wheel-base, 7 ft. ; cylinders, 9 ins. X 14 ins. ; 
gauge of track, 30 ins. The locomotive carries two tanks, 18 ft. 
6 ins. and 15 ft. 6ins. X 3oins. diameter, with a capacity of 160 cu. 
ft. of air at 600 lbs. Pipe-line, 4,100 ft. long; pressure, 700 lbs. 
Total charging time, i min. 25 sees. After reduction to 125 lbs. 
working pressure the air is reheated. Trains usually consist of 
30 cars, each weighing loaded, 5,850 lbs., though the locomotive 
has a capacity of 50 cars. Grades, from 9 ins. per 100 ft. against 
the load, to 12 ins. per 100 ft. in favor of the load. Round-trip 
time, for 8,200 ft. plus a switching distance of 800 ft., 16 min. 
Cost of haulage per ton-mile, gross, about i J cents. 

6. The Anaconda Copper Mine, Butte, Mont., is provided 
with a number of compressed-air locomotives with 5-in.Xio-in. 
cylinders and weighing 10,000 lbs. Over all dimensions: 
height, 58 ins. ; width, 58 ins. ; length, 10 ft. 4 J ins. ; four driving 
wheels, 23 ins. diam. ; wheel-base, 36 ins., designed for curves of 
i2-ft. radius. Capacity of main tank, 47 cu. ft.; pressure, 550 lbs. 
working pressure, 125 lbs.; charging time, 60 sees. Length of 
haul, 2,400 ft. round trip; load, 6 cars, weighing loaded 3,450 lbs. 



484 COMPRESSED AIR PLANT 

each ; track nearly level. The locomotives are designed to make 
2 round trips, or 4,800 ft. on i charge, with cold air; but, by re- 
heating v^ith hot v^ater, 3 round trips can be made. 

At the nev7 reduction works of the Anaconda Company, there 
are 13 H.K. Porter locomotives, employed in handling the prod- 
ucts between the different divisions of the plant, which covers 
roughly an area of 2,200X2,300 ft., the length of haul ranging from 
1,000 to 7,000 ft. Twelve of the locomotives have the following 
dimensions; weight, 26,000 lbs.; cylinders, 9 J X14 ins,; driving 
wheels, 28 ins.; wheel-base, 54 ins.; main tanks, 132 cu. ft.; 
draw-bar capacity, 5,700 lbs. Another locomotive weighs 
42,000 lbs.; cylinders 12X18 ins.; driving wheels, 36 ins.; wheel- 
base, 60 ins.; main tanks, 218 cu. ft.; draw-bar pull, 9,180 lbs. 
Tank pressure, 700 to 800 lbs.; working pressure, 150 lbs.* 

7. The Homestake Mining Co., Lead, S. D., employ under- 
ground 10 H. K. Porter locomotives, weighing 9,500 lbs. and 
measuring over all, 4 ft. 11 ins. high X3 ft. 3 J ins. wide X 10 ft. 6 
ins. long. Gauge of track, i8ins. They have a detachable rear end 
(similar to those of the Loretto Iron Co., mentioned in the 5th 
column of Table LVIII) to permit of transferring them from 
level to level, on a cage with a 9-ft. platform. At the same mine 
a small locomotive, with 5 X 8-in. cylinders (see Table LVIII) 
has been recently installed. This size is found more satisfac- 
tory, for the underground conditions prevailing in the mine, than 
the larger locomotive, with 6Xio-in. cylinders. 

8. Several 4-cylinder, Vauclain compound air locomotives, 
built by the Baldwin Locomotive Works, are in use in one of the 
collieries of the P. & R. C. & 1. Co.f Cylinders 5 and 8 ins. X 12 
ins. stroke, with valves of the balanced-piston type. Tank press- 
ure, 600 lbs. ; working pressure, 200 lbs. Driving wheels, 24 
ins. ; wheel-base, 48 ins. ; storage tanks, of j-q in. plate, 1 1 f t. 4 J ins. 
and 13 ft. 7 J ins. X 31 ins. diameter; auxiliary reservoir, 8 ins. 
diam. X 7 ft. 4 ins. long. Over all dimensions : 6 ft. 4 ins. wide X 

* A detailed description of this haulage plant is given by C. B. Hodges, Gassier' S 
Magazine, 1905. 

t Engineering and Mining Journal, Aug. 19th, 1899, p. 218. 



COMPRESSED AIR HAULAGE 485 

14 ft. long X 6 ft. 6 ins. high; weight, 22,000 lbs. Trains of 32 
cars, each weighing loaded about 9,000 lbs., are hauled on i|- 
per cent, grade, in favor of the load. 

9. At the Aragon Iron Mine, Norway, Mich., is an H. K. 
Porter locomotive. Weight, 7 tons; height, 5 ft. 2 ins.; width, 

4 ft. 2 ins.; length, 12 ft., over all. Four 24-in. drivers; wheel- 
base, 48 ins. ; gauge, 22 J ins. ; cylinders, 7X12 ins. ; tank pressure, 
700 lbs. ; working pressure, 140 lbs. ; charging time, 30 to 60 sees. 
Haulage distance, from 1,200 to 4,000 ft. ; pipe-line, 1,800 ft. ; in- 
cluding 750 ft. down the shaft. Locomotive hauls four 20-car 
trains per 10 hrs., from each of 10 loading places. Weight of 
loaded train, including locomotive, 43 tons; weight empty train, 18 
tons. Compressed air is furnished by a Norwalk 3-stage charger: 
steam cylinders, 14X16 ins.; air cylinders, loj, 7|,and 2|ins. X 
16 ins., compressing 125 cu. ft. free air per minute to 800 lbs. 
At the compressor there are two receiver storage tanks, each 
3x17ft. 

10. Compressed-air haulage plant at No. 6 Colliery of the 
Susquehanna Coal Co., at Glen Lyon, Penn. Following is an 
abstract of tests made by J. H. Bowden and R. V. Norris.* 
Though the plant is not of the latest pattern, the results given will 
be found useful. 

Compressor: Norwalk, 3-stage; steam cylinder, 20X24 ins.; 
air cylinders, 12 J, 9J, and 5 ins. X 24 ins.; capacity, at 100 revo- 
lutions, 296 cu. ft. free air per minute, compressed to 600 lbs. 
Main pipe-line at No. 6 shaft, 4,380 ft. long, 5 ins. diameter, with 

5 charging stations, and capacity of 608 cu. ft. Branch line, in 
No. 6 slope, 3,100 ft. long, 3 ins. diam., with 3 charging stations, 
and capacity of 159 cu. ft. 

Locomotives : two, by H. K. Porter Co. ; weight, 8 tons ; tank 
capacity, 130 cu. ft.; pressure, 550 lbs. reduced to 160 lbs. in an 
8-in. auxiliary reservoir, of 4.2 cu. ft. capacity. Cylinders, 7 X 14 
ins. ; four 24-in. drivers. 

At No. 6 shaft the run averages 4,000 ft. each way, on ^ to 2f- 
per cent, grades, averaging about i per cent, in favor of the load. 

* Transactions American Institute of Mining Engineers, Vol. XXX, p. 566. 



486 



COMPRESSED AIR PLANT 



Run at No. 6 slope averages 2,100 ft., with nearly the same 
grades. Cars weigh 2,800 lbs. empty, and about 9,800 lbs. 
loaded, and are hauled in trains of 12 to 20. The shaft loco- 
motive hauls about 355, and the slope locomotive 320 cars, per 10 
hours, doing the work of 32 mules. Tests on the compressor 
showed 150 indicated horse-power at 131 revolutions, compress- 
ing 387.8 cu. ft. free air per minute. 

The combined capacity of both pipe-lines is 767 cu. ft., 
which, at 600 lbs. gauge pressure, is equivalent to 32,500 cu. ft. 
free air. Capacity of locomotive main and auxiliary tanks, 134.6 
cu. ft. At 508 lbs. (at which pressure the tanks equalize with the 
mains, the initial pressure being 600 lbs.), this is equivalent to 
4,845 cu. ft. free air. In standing 12 hours, the pipe-line pressure 
falls to 350 lbs., the loss per hour from leakage thus being 974 cu. 
ft. free air, or 4.18 per cent, of total air compressed. 

Table LX 

Air Consumption 



Shaft Loco, 



No. 2 
Plane. 



No. -i 
Plane. 



Slope 
Loco. 



Number of trips, empty 

Number of trips, loaded 

Average number cars per trip, empty 

Average number cars per trip, loaded 

Average cu. ft. free air per trip, empty 

Average cu. ft. free air per trip, loaded 

Average cu. ft. free air per round trip 

Average cu. ft. free air per ton-mile, on gross tonnage 
Average cu. ft. free air per ton-mile, on net tonnage . 



3 

3 
15-33 

13 
1,724 
1,631 
3,355 



10 

10 

12.7 

13 
5,686 
1,898 
7,584 



113 
203 



16 

15 
II. 4 

11-3 
1,230 

599 
1,829 



71 

128 



Average volume free air used by both locomotives per ton- 
mile was: gross, 100 cu. ft. ; net, 180 cu. ft. The greater quantity 
of air used by the shaft locomotives is due to the heavier grades and 
switching required. Another test showed a total consumption of 
223,020 cu. ft. free air, for hauling 676 cars per day. The volume 
of free air apparently compressed for this work was 279,200 cu. 



COMPRESSED AIR HAULAGE 487 

ft., of which 83.4 per cent, is accounted for, leaving 16.6 per 
cent, for leakage and slip in the compressor, leakage in air lines, 
and changes in temperature. 

The cost of the plant, omittingboilers, was : 

Compressor and extras $2,955 • 75 

Two locomotives and extras 5,869 . 76 

Pipe-line: 5-in. line, 6,000 ft $2,914.32 

3-in. line, 4,000 ft 1,240.46 4,154.78 

Steam connections to compressor 278. 27 

Material and labor, compressor house and foundations, installing pipe- 
line, etc 1,525.23 

Charging stations 372 .21 

Total cost $15,156.00 

The average cost of operation of entire plant, for 2 years, on 
basis of 170 days' work per year, was $12.60 per lo-hour shift, 
including an allowance of $2.32 for steam for compressor, fur- 
nished by main boiler plant of mine. Adding proportion of 
fixed charges, with interest, depreciation and repairs, the daily 
cost, on basis of 300 days' work per year, would be $18.52 per 
day. For the 2 years, the average cost per ton-mile was as follows : 

Table LXI 

Operating CostS^ 





1897 (179 Days). 


1898 (160 Days). 






1 




f 


to 



"a 

Q 




Shaft locomotive, gross tonnage 

Shaft locomotive, net tonnage 

Slope locomotive, gross tonnage 

Slope locomotive, net tonnage 

Both locomotives, gross tonnage 

Both locomotives, net tonnage 


1,485 

648 
360 

2,133 
1,185 


$11.12 
II. 12 
II. 12 
II . 12 
22.23 
22.23 


0-75 

1-35 
1.72 

3-09 
1.05 
1.89 


1,521 

845 

720 

400 

2,241 

1,245 


$12.00 
12.00 
12.00 
12.00 
24.01 
24.01 


0.79 
1.42 

1.67 

3.00 
1.07 

1-93 



In these two years the saving over the expense of the mule 



488 COMPRESSED AIR PLANT 

haulage, previously employed, was $14,218.00, or nearly the total 
cost of the haulage plant. 

II. Following is the cost of a large colliery plant, as given by 
Beverly S. Randolph,* who installed and afterward operated it: 

Three-stage, compound compressor $5,300. 

Pipe line: 5,600 ft., 5 in $5,600. 

3,100 ft., 2 J in 1,700. 

1,000 ft., i| in 300. 

— — 7,600. 

Two main locomotives, weight 30,000 lbs 6,000. 

Five gathering-locomotives, weight 8,000 lbs 10,000. 

Two boilers, each 80-horse-power : 1,000. 

Installation, labor, and material 4,000 . 

Total cost $33,900 . 

This plant includes an unusually large number of small 
gathering-locomotives, for collecting cars from the individual 
workings and making them up into trains for the main haulage 
lines. If the locomotive equipment had consisted of four2 5,ooo-lb. 
engines, costing, say, $2,800 each, and which would do the same 
work, the total cost of the plant would be reduced to $29,100. 
This cost compares very favorably with that of electric-haulage 
plants of the same capacity. 

'^ Transactions Institution 0} Mining Engineers (England), Vol. XXVII (1904)^ 
P- 433- 



INDEX 



Abrams, H. T., test on air-lift pumps, 450 

Absolute pressure, temperature and zero, 49 

Adelaide drill, 300 

Adiabatic compression, 51-56, 58-66, 68-74, 113, 160-164; equation of, 60, 68 

Adiabatic expansion, 265-267 

Adjustable steam cut-off valve, 34 

Aftercooler, no, 192 

Ainsworth (B. C), hydraulic air compressor at, 241, 242 

Air and steam cards combined, 35 

Air card, 57, 61, 64, 67, 70, 72, 112, 172, 173, 214; elements of, 168; of wet and 
dry compressors, 63; of stage compressor, 112, 172 

Air-cataract valves, 139 

Air compression: at altitudes above sea-level, 216-222; by direct action of falling 
water, 235-247, 292 

Air compressors: belt-driven, 12, 36, 44-46; builders, list of, 48; chain-driven, 
12, 44; classification of, 8, 9, 10, 12; dry, 62, 63, 81; for compressed-air haul- 
age, 474-479; geared, 44, 46; half-duplex, 16; horse-power of, 160 et seq.; 
hydraulic, 235, 247, 292; makers of (see Compressors); performance of, 
159-189; water-driven, 36-44; wet, 62, 63, 75 

"Air-Electric" drill, 321 

"Air-Electric" track channeler, 418, 421 

Air engines, 261-272, 282, 284, 290 

Air-feed hammer drills, 343, 35°. 352, 357, 360, 366, 369 

Air governors, 196-215 

Air hammer drills, 348 et seq. 

Air inlet, area of, 117, 118, 127; conduit for, 134 

Air inlet valves, 115-135, 142-158 

Air-lift pumps, 438 et seq. 

Air pressure for machine drills, 324, 350, 358, 363, 366, 377 

Air pressure regulators, 196 et seq.; American, 198; Clayton, 197, 198, 201; 
Ingersoll-Sergeant, 206; Laidlaw-Dunn-Gordon, 207; Nordberg, 209-215; 
Norwalk, 198-200; Rand, 203; Sullivan, 204, 205 

Air pulsator for "electric-air" drill, 321 

Air receivers, 150 et seq.; functions of, 191; sizes, 190; underground receivers, 
192, 193, 277, 432 

Allis-Chalmers compressors, 12, 29, 30, 116 

AlHs-Chalmers mechanically controlled valve-motions, 149 

Altitudes above sea-level, compression at, 216 et seq. 

American Air Compressor Works, 48 

American air-pressure regulator, 198 

American Institute of Mining Engineers, Transactions of, 229, 232, 485, 488 

American Locomotive Co. compressed-air locomotives, 462 

American Machinist, 112, 209, 217, 255, 272, 291 

Anaconda Copper Mine, compressed-air haulage at, 483, 484 

Angelo and Cason Mills, South Africa, tests on air-lift pumps, 450-452 

Anthracite Coal Operators' Association, Transactions of, 468 

Aragon Iron Mine, Mich., compressed-air haulage at, 485 

489 



490 INDEX 

"Arc-valve" tappet drill, 311, 312, 313 

Area of air inlet, 117, 118, 127 

Auger coal drills, 407 

Area of discharge valves, 140 

Association of Engineering Societies, Transactions of, 435 

Auxiliary reservoirs for compressed-air locomotives, 466, 468 



"Badger" drill, 309 

Baffle plates for air receivers, 107, 195 

Bailey & Co., Manchester, piston valve, 158 

Baldwin Locomotive Works compressed-air locomotives, 460, 462, 465, 482^ 484 

Ball-and-socket joints for compressed-air locomotive charging station, 470-472 

Barre quarries, Vermont, 374 

Barrow drill, 300 

Behr, H. C, 273; air-lift pump experiments, 447 

Bell, J. E., experiments by, 328 

Belt-driven compressors, 12, 36. 44-56 

Bendigo district, Victoria, Lansell's air-lift pump, 453 

Bends in air pipe, 260 

Bernstein, 246 

Bjorling, P. R., 76, 90 

Bleeder valve for compressed-air locomotive charging-station, 472 

Bowden, J. H., test on compressed-air haulage plant, 485-487 

Boyer hammer drill, 371 

Boyle's law, 50 

Breakage of drill parts, 333 

Buck Mountain Colliery, compressed-air haulage at, 479 

Burleigh compressor, 2 

Burning-point of cylinder oils, 225, 227 

Burra-Burra Mine, compressor at, 20; drills at, 375 

Butte, Montana, compressor explosion, 282 

By-pass for air cyhnder, 90 



Cable-reel electric locomotive, 457 ^ 

Calumet and Hecla Copper Mine, 3 

Cam-controlled poppet inlet valve, 156 

Canadian Electrical News, 241 

Canadian Engineer, The, 236 

Canadian Mining Institute, Transactions of, 277 

Capacitv of air for moisture, 2 74 

Cards, air, 57, 61, 63, 64, 67, 70, 72, 112, 168, 172, 173, 214 

Carnahan, C. T., Manufacturing Co., 371 

Carper, J. B., experiments by, 324 

Cason and Angelo Mills, South Africa, tests on air-lift pumps, 450-452 

Cassier^s Magazine, 484 

Cataract valves, 138 

Causes of freezing of moisture in compressed air, 275 

Cave rock drill, i 

Chain-driven compressor, 12, 44 

Champion Iron Mine, experiments at, 330 

Channehng machines, 409 ct seq.: depth of cut and speed of work, 418; shape 

of bits, 414, 417 
Channing, J. Parke, 20, 170 

Charging compressor for compressed-air locomotives, 474-479 
Charging station? for compressed-air locomotives, 470-472 
Charles's law, 51 



INDEX 491 

Chattering of inlet valves, 119, 120 

Chersen drill, 300, 378 

Chicago Pneumatic Tool Co., 371 

Chodzko, A. E., 273 

Choking of air pipes by ice, 275, 277, 278 

Christensen compressor, 46, 48 

"Cincinnati" air-valve gear, 139, 14S 

Clack valves, 11 s, 131 

Clark, D. K., 51^ 

Classification of compressors, 8, 9, 10, 12 

Clausthal Silver mines, hydraulic compressor at, 246 

Clayton compressor. 2, 48; governor for, 197, 198, 201 

Clearance: in compressor cylinder, 66-74, 85-90, 113, 116, 221; in air engine, 
265, 268-270 

Clearance: proportionate and disproportionate, theory of, 70-74 

Clemens, G., 482 

Cleveland hammer drill, 371 

Cleveland Pneumatic Tool Co., 371 

Clifton Colliery, England, explosion in compressor, 226, 233 

Climax "Imperial" drill, 300, 315 

Climax hammer drill, 369, 371 

Coal cutting machines, 380 et seq.; comparison of, 407; disc or circular saw, 
385; depth and width of cut, 384, 385, 389, 394, 397; endless-chain, 380- 
385; pick or reciprocating, 381, 388-405; rotary-bar, 380, 385 

Coal punchers, 388-405 

Colladon, i, 78 

Colliery Guardian, 277 

Colorado Fuel Co., 406, 407 

Comparison of types of compressors, 18 

Complete expansion, working with, 264-267 

Compound compressed-air locomotives, 474-484 

Compound steam-end for compressors, 22, 24-32 

Compressed-air drills, 294-379 

Compressed-air engines, 261-273 

Compressed-air haulage, 456-488 

Compressed Air Machinery Co., 36, 48 

Compressed Air Magazine, 141, 195, 229, 245, 270, 272, 427, 441, 450, 482 

Compressed-air pumps, 277, 423-437, 438-455; adjustment of air pressure, 432; 
efficiencies of, 431-433; preheating for, 434-437; prevention of freezing 
in, 277, 433 

Compressed air, reheating of, 279-293 

Compressed air versus electric transmission, 5,6; versus steam transmission, 3, 4, 5 

Compressed air versus steam for direct-acting pumps, 425-427 

Compression curve, construction of, 166 

Compression of air: laws, 50 et seq.; at altitudes above sea-level, 216-222; by direct 
action of falling water, 235-247; heat of, 53 et seq.; stage compression, 63 et 
seq.; 95-114, 160-164 

Compressors, makers and names of: Allis-Chalmers, 12, 29, 48, 116; Burleigh, 
2; Chicago Pneumatic Tool Co., 48; Christensen, 46, 48; Clayton, 2, 48; 
De la Vergne, 34; Dubois-Franfois, 2, 95 115; Franklin, 48, 152, 198; 
Hanarte, 77; Humboldt, 75, 76, 130, 139; Ingersoll-Rand, 9, 12, 13, 17, 25, 
26, 36, 41, 45-48, 8 , 198, 474, 476, 478, 479; Johnson, 89, 129; King-Riedler, 
10, 16; Laidlaw-Dunn-Gordon, 9, 10, 11, 12, 14, 15, 81, 83, 84, 86, 116, 
139, 147, 148; Leyner, 12, 21, 26, 48, 107, 131, 132, 133; Nordberg, 19, 48, 
81, 82, u6, 209-215; Norwalk, 2, 12, 19, 20, 48, 99, 100, 116, 125, 474, 483, 
485; Rand, ?, 3, 85, 115; Rand and Waring, 34; Riedler, 12, 27, 28, 139; 
SulHvan, 12, 22, ^3, 48, 105, 106, 150, 151, 475 

Congelation of moisture in compressed air, 93, 274-278, 433 

Consumption of air: by air engines, 270, 271, 283-284, 302; by direct-acting 



492 INDEX 

pumps, 428-431; by machine drills, 324-330> 352-35^, 360, 366,377; by 

pneumatic pumps, 440, 441, 447, 449, 450, 452 
Conveyance of compressed air in pipes, 248-260 
Cooling: modes of, 55, 63-65; in receivers, 191, 194, 276, 277 
Corliss air valves, 43, 142-152, 209-213 
Corliss steam-valve motion for compressors, 20, 22, 29-31 
Couch, J. J., machine drill, i 
Cox, Wm., 198, 252, 427 
Crane, W. R., 340 
Cresson Mine, Cripple Creek, 341 

Cummings, Chas., system of compressed-air transmission, 272, 437, 442 
Cushioning in machine drills, 303, 311, 312, 332, ^^;^ 
Cut-off in air engines, 266-271; nominal and actual, 268-270 
Cylinder dimensions of simple pumps, 427 
Cylinder proportions for compressors, 35 
Cylinder volumes: in stage compression, 101-104; of air engine, 270 



''Dancing" of inlet valves, 119, 120 

D'Arcy formula for loss of pressure in pipes, 251-255 

Darhngton drill, 300, 319 

De Kalb, 111., tests on air-lift pump, 447 

Delivery valves, 136-141; cataract-controlled poppets, 138; effect of leakage, 136; 

mechanical control for, 143; poppet type, 136, 146, 148, 149, 150, 152 
Denver Rock Drill and Machinery Co., 371 
Deposition of moisture from compressed air, 274-277 
Dickson Manufacturing Co. compressed-air locomotives, 483 
Dingier Machine Works, Zweibruecken, 131 
Dinnendahl, R. W., 319 

Direct-acting pumps, operation by compressed air, 423-437 
Direct compression by falling water, 235-247 
Disc or circular saw coal cutters, 380, 385-388 
Discharge valves (see Delivery valves) 
Discharge-valve area, 140 

Displacement pumps, pneumatic, 438-443; consumption of air by, 440, 441 
Doble drill, 300 

Dover Iron Co., compressor, 158 
Drill repairs, 333 

Drilling records, 334-341, 372-376 
Drills, rock: air pressure for, 36, 324, 327, 335 et seq.; 375, 377, 378; reheaters 

for, 290; repairs, 2)33) records of work, 334-341, 372-376; valve motion of, 
.332 
Drinker, Tunneling, Explosive Compounds, and Rock Drills, 294 
Drummond ColHery, compressed-air pumps at, 277 
Dry compressors, 62, 63, 81-94 
"Dry" reheaters, 289, 293 
Dry versus wet compression, 90 
Dubois-Franfois compressor, 2, 75, 115 
Duisburger Maschinenbau, 319 
Duplex compressors, 9, 10, 12, 17, 18, 23-32 
Dust allayer for machine drills, 316, 370, 378 
Duty of machine drills, 325, 329, 334-341, 371-376, 31^ 

E 

East Rand Proprietary Mines, Ltd., tests on air-lift pumps, 450-453 
Ebervale, Luzerne Co., Pa., tunnel at, 257 
Efficiencies of air-Hft pumps, 441, 447, 448, 449, 452 
Efficiencies of direct-acting compressed-air pumps, 431-433 



INDEX 493 

Efficiency of compressors, 20, 108, no, 122, 133, 171, 172, 176-189, 240, 244 

Efficiency of reheating, 281-284 

Electric-driven compressor, 44, 46, 47 

"Electric-air" drill (see "Air-electric" drill) 

"Electric-air" track channeler, 418, 421 

Electric versus compressed-air haulage for mines, 456-458, 488 

Electric versus steam locomotive haulage for mines, 456 

Empire Mine, Grass Valley, Cal., compressed-air haulage at, 480 

Endless-chain coal cutters, 380, 385 

Engineer, The (London), 450 

Engineering and Mining Journal, 236, 240, 244, 246, 330, 436, 484 

Engineering News, 89, 375, 447 

Esmeralda Mine, Silverton, 375 

Expansion curves, air and steam, 263, 267 

Explosions in air compressors and receivers, 223-234 

Externally heated or "dry" reheaters, 289, 293 

F 

Fergie, Chas., 227 

Ferroux drill, 300 

Final temperature of air compression, 53, 54, 165, 166, 169, 170, 224 et seq. 

"Fitchering" of drill holes, t,t,t„ 334, 340, 341, 348 

Flash and ignition points of cylinder oils, 225, 227 

Flat River, Mo., drilling records in lead mines, 337 

Flottmann & Co., 371 

Flottmann hammer drill, 371 

Four-stage compressor, 474-478 

Fowle machine drill, i 

Franke hammer drill, 343, 371 

Franklin compressor, 48, 152, 198 

Franklin pressure regulator and unloader, 198, 206 

"Free" air, 49 

Freezing of moisture in compressed air, 93, 274-278, 433, 434 

Freimann & Wolf, 319 

Frick, H. C, Coal and Coke Co., 406 

Frictional losses in compressors, 20, 97, no, 159, 162, 171, 175 et seq., 186 

Frictional resistance in air pipes, 248-260; due to bends, 260 

Friedrich, G. C. H., tests on air-lift pumps, 449 

Frizell, J. P., 236 

Froelich & Kliipfel, 319 

Fuel cost of reheating, 283-285, 293 

Full pressure in air engines, working with, 265 

Functions of air receiver, 190-194 

G 

Gay-Lussac's law, 51 

Geared compressors, 44, 46, 47 

General Electric Co., 385 

Gernaan machine drills, 300, 319 

Gillott and Copley coal cutter, 388 

Glen Lyon, Pa., Colliery, compressed-air haulage at, 485-487 

Gliickauj, 246 

Goleta Mining Co. water-driven compressor, 36 

Gordon hammer drill, 378 

Governors, air, 196-215; American, 198; Clayton, 197, 198, 201; Franklin, 
206; Ingersoli-Rand, 198, 201, 204, 207; Ingersoll-Sergeant, 206, 207; Laid- 
law-Dunn-Gordon, 207, 208; Nordberg, 207, 209-215; Norwalk, 198-200; 
Rand, 202; Sullivan, 198, 204, 205 

Grades of mine haulage tracks, 472, 479-486 



494 INDEX 

Great Western Pneumatic Tool Co., 371 
Guttermuth air valve, 115, 131 
Guttermuth, experiments on reheating, 284 
Gwin Mine, Cal., pump reheater, 436 

H 

Halsey, F. A., 112, 217, 218, 255, 256 

Halsey pneumatic displacement pump, 443 

Hammer drills, air, 343-379; air pressure for, 377; depth of hole and speeds of 
work, 371-376; makers of, 371 

Hanarte compressor, 77 

Hardsocg hammer drills, 348, 374 

Hardsocg Wonder Drill Co., 371 

Harris pneumatic displacement pump, 442, 443 

Harrison pick machine, 388, 392-395 

Haulage by compressed-air locomotives, 456-488 

Heat curves, 54 

Heat losses in compressors, 91 

Heat of compression, 51 e/ seq., 78, 79, 84, 91, 95 

Heat, transference of, 55 

Heating of air-cylinder walls, 114 

Henderson, tests on air-lift pumps, 450-453 

Heron & Bury Manufacturing Co., 48 

Highland Boy Mine, Bingham, Utah, drilling record, 339 

High-pressure transmission of air, as influencing freezing, 276 

"High-range" compressed-air transmission, 272, 273, 437, 442 

Hill, E., 229, 230 

Hirnant drill, 300 

Hiscox, G. D., 268 

Hodges, C. B., 484 

Hoffman, P., 319 

Holman drill, 300, 316-318 

Homestake Gold Mine, compressed-air locomotives at, 458, 484 

Hoosac tunnel, 2, 256 

Horse-power: of air engines, 265-271; of air-lift pump, 448-450, 452; of com- 
pressors, no, 159 et seq., 240, 244 

Horse-power per cu. ft. of free air, 160-164 

Humboldt compressor, 75, 76, 130, 139; rubber-ring valve, 130 

Humboldt Machine Works air-cataract valve, 139 

Humidity of atmospheric air, 93, 274, 275 

" Hurricane inlet " valve, 126-129 

Hydraulic air compressor, 235-247, 292 



Ignition-points of cylinder oils, 225, 227 

Her hammer drill, 371 

Her Rock Drill Manufacturing Co., 371 

Ingersoll-Rand Co., "air-electric" drill, 321; channelers, 411, 413, 417, 419; 

compressors, 9, 12, 13, 25, 31, 36, 37, 41, 42, 45, 47, 86, 112, 118, 476-478; 

drills, 300-304, 311, 321, 335, 336, 338, 341; hammer drills, 344, 358-366; 

intercooler, 105, no, in; pick machine, 394, 395-397; "Radialaxe" coal 

cutter, 400; ram track channeler, 413, 421; receiver, 192; steam regulator, 

201, 206 
Ingersoll-Sergeant piston-inlet valve, 116, 126-129 
Injection water: effects of, on air cylinder, 91; quantity of, 79; temperature of, 

78-80 
Inlet air, arrangements for admitting, 134 
Inlet valves, 115-134; area of, 117, 118, 127; chattering of, 119, 120 



INDEX 495 

Institution of Civil Engineers (London), Proceedings of, 236 

Intercoolers, 96, 100 et seq.; Ingersoll-Rand, 105, no, in; Leyner, 109; Norwalk, 

100; Schram, 108; Sullivan, 105, 106; volume of, 102, in 
Internally fired reheaters, 288 
Isothermal compression, 55, 56, 61, 63, 67, 161, 164 



Jeddo (Pa.) Mining tunnel, compressed-air transmission in, 257 

Jeffrey Mfg. Co., 309, 385 

Jeffrey "Badger" drill, 300, 309 

Jeffrey coal cutters, 380, 382-388 

Johannesburg stope drill contest, 377 

Johnson compressor, 89; air valves of, 129 

Johnson, E. E., on performance of air-lift pump, 477, 478 

Joule's heat unit, 53, 168, 169 

K 

Kennedy, Alex. B. W., reheating tests by, 383, 393 

Kimber hammer drill, 343, 371 

King-Riedler compressor, 10, 16 

Knight water-wheel, 36 

Knowles Steam Pump Works, 48 

Konomax drill, 300 

Kootenay (B. C), hydraulic air compressor at, 241, 242 

Koster piston air valve, 158 

Kiizel drill, 300 

L 

Laidlaw-Dunn-Gordon Co.: compressor, 9, 10, 11, 12, 14, 15, 48, 81, 83, 84, 86, 
116, 118, 119, 139, 147-149; "Cincinnati" valve gear, 139, 148; mechanically 
controlled valve motions, 147-149; poppet inlet valve, 119; pressure regulators, 
207, 208 

Lansell's air-lift pump for shafts, 453-455 

Latta-Martin displacement pump, 441 

Laws governing compression of air, 50 e/ seq. 

Leakage of compressed-air pipe lines, 252, 256, 259, 486, 487 

Leaky air piston, effect of, 113, 229, 231 

Leaky discharge valves, 223, 229, 231 

Lees, T. G., 85, 224, 226 

Lens Colliery, France, compressor air valve motion at, 156 

Leyner, J. Geo., Eng. Wks. Co., 48, 371 

Leyner compressor, 12, 21, 26, 48, 131-133, 134; intercooler, 107-109; reheater, 
285, 286 

Leyner drills, 343, 344, 345-348, 37^, 372, 373, 375 

Lightner Mine (Cal.), reheating at, 292 

Link Belt Machinery Co., 385 

Lippincott, J. B., 375 

"Little Champion" drill, 312 

"Little Jap" drill, 371 

"Little Wonder" drill, 371 

Locomotives, compressed air, 456-488; construction and operation, 458 et seq.; 
cylinder pressures, 466, 469 

Locomotive charging compressors, 474-479 

Long-wall coal cutters, 384, 385-388 

Lord, E. P., 468 

Los Angeles aqueduct tunnel, 375 

Loss of head in pipe transmission, 249-258; of power, 248 

Loss of volumetric capacity due to piston clearance, 85, 86, 88, 89 



496 



INDEX 



Losses in air compression, 159-162 

Losses in transmission piping, 248-260 

Lubrication of air cylinders, 93, 223, 225, 227-229, 233 

Lubricators, sight-feed, 233 

M 

Machine drills, 294 f/ seq.; classification, 299; column or bar for, 299; consump- 
tion of air by, 324-330, 352, 358, 360, 377; cushioning of stroke, 303, 311, 312, 
332; dust allayer for, 316, 370, 378; "electric-air" drill, 300, 321; feed of, 
295; general description, 295; hammer drills, 294, 343-379; length of stroke, 
295, 296; makers of, 300, 371; modes of mounting, 296; records of work, 
334-341, 372-396; repairs of, 2)2>2>\ rotation of bit, 295, 303, 304, 321; screw 
feed, 295; sizes of, (see different makes of drill); speeds of drilling, 325, 
335~34i) 37I; 372-376; speed of stroke, 295, 350, 363, 366; spool-valve drills, 
3oo> 304, 309^ 315- 317; tappet-valve drills, 309, 311, 312, 317 

Machine drills (hammer), makes of: Boyer, 371; Cleveland, 371; Climax "Im- 
perial," 369; Flottmann, 371; Franke, 343, 371; Her, 371; Ingersoll-Rand 
"Crown," 358; Kimber, 343, 371; Leyner, 345; "Little Jap," 371; Murphy, 
351; Ingersoll-Rand "Imperial," 363; Schmucker, 371; Shaw Eclipse, 371; 
Sullivan, 354; Waugh, 366; Whitcomb, 371; Wonder, 348 

Machine drills (reciprocating), makes of: Adelaide, 300; Barrow, 300; Chersen, 
300, 378; Climax "Imperial," 315; Darlington, 300, 319; Doble, 300; Ferroux, 
300; Froelich, 300; Hirnant, 300; Holman, 317; IngersoU, 300,311; Jeffrey, 
309; Konomax, 300; Kiizel, 300, 319; "Little Wonder," 300; McKiernan, 
300; Meyer, 300, 319; Murphy, 312; Rand, 300; Rio Tinto, 300; Rix, 300; 
Schram, 300; Sergeant, 300, 303; Sierra, 300; Sullivan, 304, 309; Temple- 
Ingersoll, 321; Triumph, 319; "WahrwoP," 300; Wood, 300 

Magog (Prov. Quebec), hydraulic air compressor, 236-240, 241 

Mallard, M., 264 

Mannesmann tubes for high air pressures, 465 

Mansfeld copper mines, underground receivers, 193 

McKiernan Drill Co., 48; air-pressure regulator, 198; drill, 300 

McLeod, C. H., 240 

Mean pressures in air compression, 165, 169 

Mechanical Engineers' Assocn. oj the Witwatersrand, Trans, of, 273, 324 

Mechanically controlled air valves, 115, 117, 142-158; Allis-Chalmers, 149; 
American, 152; Clayton, 152; disadvantages for delivery valves, 143, 149; 
for high altitudes, 220; Franklin, 152; Laidlaw-Dunn-Gordon, 148; Nord- 
berg, 144, 147; Norwalk, 144, 145; Riedler, 152-156; Rix, 152; Sullivan, 
150, 151 

Menck and Hambrock cataract valve, 139 

Merrill pneumatic pump, 439, 440 

Meyer steam valve gear, 9 

Meyer, R., air-cataract valve, 139; machine drill, 300, 319 

Michigan Copper Mine, drilling record, 336, 373 

Midlothian colliery, Va., 374 

Mines and Minerals, 20, 244, 324, 405, 479 

Mining and Scientific Press, 334 

Mining and Metallurgy, 328 

Modern Machinery, 273 

Moist air, effect of, in compression, 91, 92, 93 

Moisture in air, 84, 91-93, 191, 193, 195, 274-278 

Mont Cenis tunnel, 1,2; speed of advance in, 2 

Morning Mine, MuUan, Idaho, compressor plant at, 40, 42-44 

Mount Hope Iron Mine, drilling record, 337 

Mule haulage, cost of, 480, 483, 486 

Murphy hammer drill, 351-354, 37 1, 373, 375 

Murphy reciprocating drill, 312, 314 

Mushroom valve, 118, 119, 124 



I 



INDEX 497 



N 



"n " values of, 52, 62, 84, 113, 160, 163, 164, 168, 224 

Neumann and Esser piston air valves, 158 

New IngersoU pick machine, 395 

New Reitfontein, mine, S. A., 374 

New York Air Compressor Co., 48 

New York x\queduct, explosion in compressor, 229 

Nicholson, J. T., experiments by, 292 

Nominal and actual cut-offs in air engines, 268-270 

Non-conducting covering for air pipe, 290, 291 

Nordberg compressor, 20, 48, 81, 82, 116, 144, 147, 207, 209; air-pressure regula- 
tor, 207, 209-215 

Norris, R. V., tests on compressed-air haulage plant, 485-487 

North Star Mine (Cal.), compressor at, 40-42, 44; reheating at, 291 

Norwalk compressor, 2, 12, 19, 48, 99, 100, 116, 118, 125, 474, 475, 480, 483, 485; 
intercoolers, 100, 105; poppet inlet valve, 118; "skip-valve," 124; pressure 
regulators, 198, 199, 200; receiver, 191 

Norwich (Conn.), hydraulic air compressor, 245 



Ohio Society of Mech., Elec, and Steam Engineers, Trans, of, 449 

Oil-cataract delivery valves, 138, 139 

Oils, lubricating, 225-229, 233; flash and ignition points, 225, 227; oxidation of, 

62, 224, 225, 227 
Operation of compressors, 32, 34: stage, 98-104, no; belt-driven, 44 
Output of compressors, 17, 43, 44, 56-74, 89, 91, 108, no, 159-189, 219, 220, 240, 



243, 244, 246 



Paris Pneumatic Supply Co., 108, 283 

Partial or incomplete expansion in air engines, 266 

Peerless Colliery (W. Va.), compressed-air haulage at, 482 

Pelton water-wheel, 36-39, 40, 41, 42 

Pennsylvania Copper Mine, Butte, drilling records, 338 

Penoles, Compania de, compressor at, 46 

Performance of air compressors, 159-189, 219, 240, 244 

Phila. and Reading C. and I. Co., compressed-air locomotives, 474, 482, 484 

Phillips Rock Drill Co., 309 

Pick machines for coal mining, 380, 388 et seq.; sizes of, 389, 394, 397, 398, 404 

Pipe Hne for compressed-air haulage, 468; calculations for, 469 

Pipe, nominal and actual diameters, 251, 255; bends in, 260; joints in, 259; leakage, 

252, 259, 486, 487; precautions in laying, 259 
Piston air valves, 158 

Piston clearance in air-compressor cylinders, 85-89, 94 
Piston clearance in air engines, 268-271 
Piston inlet valves, Ingersoll-Rand, 116, 118, 125-129 
Piston speed of compressors, 61, 76, 77, 78, 84, 95, 112, 117, 141 
Plymouth Cordage Co., compressed-air locomotive at, 457 
Pneumatic displacement pumps, 438-443; consumption of air by, 440, 441 
Pneumatic Engineering Co., 442, 443 
Pneumelectric coal puncher, 402-405 
Pohle air-lift pump, 443 et seq. 
Pohle, Dr. JuUus, 443 

Pokorny and Wittekind piston air valve, 158 
Poppet valves, 115, 1 18-126; cam-controlled, 156; inertia of, 117, 118; sticking of, 

123 



498 INDEX 

Port area for air cylinders, 116-118, 127, 133 

Porter, H. K., Co., compressed-air locomotives, 457, 458-460, 463, 464-466, 479, 

482, 484, 485 
Portland Mine, Cripple Creek, drilling record, 341, 372 
Preheating for compressed-air pumps, 434, 436 
Prescott steam pump, 425 

Pressure of air as influencing freezing, 276, 277 
Pressure regulators, 196-215 

Prevention of freezing of moisture, 275, 276, 433, 436 
Proportions of compressor cylinders, 35, 36 
Protection from freezing of surface air piping, 278 
Pulsator for "Electric-air" drill, 321, 322 
Pumping by direct action of compressed air, 438-455 
Pumps, direct-acting, 423-427; compound pumps operated by air, 434-437; 

cylinder dimensions, 428; duty of, 429, 430; terminal and mean air pressures, 

430; volume of air consumed, 428-431 
Pumps operated by direct action of compressed air, 277, 423-427; air-lift pumps, 

443-453; pneumatic displacement pumps, 438-443 



Radialaxe coal cutter, 400-402 

Railway and Engineering Review, 236 

Rand compressor, 2, 3, 85, T15, 202; reheater, 288 

Rand drill, 3, 300, 336, 338, 339 

Rand mechanical air valve gear, 115 

Randall, B. M., experiments by, 447 

Randolph, B. S., 488 

Rating of compressors, 159, 160 

Rawhide pinions for geared compressors, 46 

Receiver after coolers, no, 192, 195 

Receivers, air, 190-195; baffle-plates for, 107, 195; functions of, 190-192; sizes of, 

190; underground receivers, 193, 277, 432 
Reciprocating coal cutters, 380, 388-405 
Reciprocating drills, 300 et seq. 

Records of work, machine drills, 334 et seq., 2,1 '2 et seq. 
Reducing valves for compressed-air pump, 432 
Reduction of pressure as influencing deposition of moisture, 277 
Re-expansion line of air card, 67, 70, 72, 167 
Regulation of compressors, 18, 20, 32, 34, 196 et seq. 
Regulators, air-pressure, 196-215 

Reheaters for compressed air, 279-293; for channelers, 411, 417 
Reheating compressed air, 279-293, 434-437, 468 
Repairs of machine drills, t^^^ 
Resistance due to bends in air piping, 260 
"Return air" system of transmission, 272, 273, 437, 442 
Richards, Frank, 99, 102, 134, 195, 255, 256, 272, 280, 291, 427 
Riedler compressor, 12, 16, 27, 28, 139; "express" discharge valve, 139, 140; 

mechanically controlled valves, 152-156 
Riedler, experiments by, 284 
Riedler steam pump, 425 

Rifle-bar for machine drills, 295, 303, 304, 305, 321 
Rigg and Meiklejohn coal cutter, 388 
Risdon water-driven compressor, 36, 39 
Rix Compressor and Drill Co., 48 

Rix Compressor, North Star Mine, 40; compressed-air locomotive, 480; drill, 300 
Rix, E. A., on compressed-air pumps, 431, 433, 435, 436, 447 
Rock-drifls (see Machine drills) 
Rockford, 111., tests on air-lift pump, 447 



INDEX 499 

"Rock Terrier" hammer drill, 348 

Rose, A. H., 244 

Rose Deep Mine, South Africa, experiments on rock-drills at, 329-330 

Rotation devices for machine drills, 295, 303, 304, 305, 321 

Rotary -bar coal cutter, 380, 385 

Rubber ring inlet valve, 130 

Ruhrthaler Maschinenfabrik, 319 

Rutland marble quarries, 418 



San Pedro Copper Mine, drilHng records, 335 

Saunders, W. L., 54, 91, 94, 141, 227, 293 

Schneider & Co., Creusot, 115 

Schram intercooler, 108 

Schram drill, 300 

Schuechtermann and Kremer oil-cataract valve, 139 

Schuetz, G. A., Wurzen, air-cataract valve, 139 

Schmucker hammer drill, 371 

Schwarz and Co., 319 

Sergeant reheater, 286, 287 

Sergeant coal pick, 389 

Sergeant drill, 300, 301, 304, 329, 338 

Seymour, L. I., experiments on rock-drills, 329 

Shaw Eclipse hammer drill, 371 

Shaw Pneumatic Tool Co., 371 

Shetucket River, Conn., hydraulic air compressor, 245 

Sibley Journal oj Engineering, 287 

Sierra drill, 300 

Simpkins, E., 388 

Single-stage compression, work of, 56-70, 160, 161, 163 

"Skip-valve" for Norwalk high-pressure air cylinder, 123, 124 

SUmes and sands pumped by air-lift, 450-453 

Snow Storm Mine, Idaho, drilling records, 336, 373 

Soap and water for cleaning air cylinders, 233 

Sommeiller, i; compressor, 2 

South African Association oj Engineers, Trans, oj, 329 

South Crofty Mine, Cornwall, drilling record, 374 

Specific heat of air: at constant pressure and volume, 55, 168 

Speer, F. W., 244 

Sperr}' coal cutter, 385 

Spiral- weld steel air pipe, 259 

Spool-valve drills, 300, 304, 309, 315, 317 

Spool versus tappet valves for machine drills, 332 

Sprague, T. W., 405 

Springs for inlet valves, 118; resistance of, 118-123, 133 

Stage compression, theory, 63 et seq. 

Stage compressors, 12, 16, 19, 20-33, 95~ii4; double-acting, 99; for high altitudes, 

221; ratio of cy Under volumes, 104; single-acting, 98; work of, 163, 164 
Stanley heading machine, 405-407 
Steam and air card combined, 35 

Steam-driven, direct-acting pumps, 423-425; cylinder dimension's of, 427, 428 
Stephens and Son, Carn Brea, 315 
Storage-battery electric locomotives, 457 
Straight-line compressors, 9, 12, 13, 17, 18-22 
Stroke, length of, for compressors, 35 
Sturgeon air inlet valve, 116, 135, 157 
Submergence of air-Uft pumps, 444-447, 4-49) 45 2 
Suction valves (see Inlet valves) 



500 INDEX 

Sullivan channeler, 410, 412, 413, 414, 417, 420 

Sullivan coal pick, 381, 391, 399 

Sullivan compressor, 12, 22, 23, 48, 105, 106, 150, 204, 475; intercooler, 105; 

reheater, 287-289; valve motions, 150, 151 
Sullivan drill, 304-308, 337, 338, 339 
Sullivan hammer drill, 344, 354-358 
Sullivan Machinery Co., 48, 385, 417, 475 
Summers, L. L., experiments by, 328 
Surface air piping, protection of, 278 

Susquehanna Coal Co., No. 6 Colliery, compressed-air haulage at, 485-487 
Sutro tunnel, Nevada, compressor at, 75 



Tailings pumps and wheels replaced by air-lift pumps, 450-453 

Tanks, capacity of, for compressed-air locomotives, 459, 460, 463, 469 et seq. 

Tappet drills, 307, 308, 309, 311, 312, 317 

Tappet versus spool valves for machine drills, 332 

Taylor, Chas. H., 236; hydraulic air compressor, 236 et seq., 292 

Technical Society of the Pacific Coast, Proceedings of, 431, 441, 447 

Temperatures employed in reheating, 281-285, 291, 293 

Temperatures of compression, 44, 53, 54, 76, 78, 79, 84, 95, 166, 225, 227, 230, 

231, 282 
Temperatures of expansion, 264, 269 
Temperatures, rate of increase of, in compression, 53 
Temple-Ingersoll "electric-air" drill, 300, 321, 400, 418, 420 
Tennessee Copper Co., compressor of, 20, 170 
Tests on compressors, 169-189, 240, 244 
Tests on machine drills, 324-326, 329, 377, 378 (foot-note) 
Theoretical horse-power required to compress air, 56 et seq., 160, 161 
Thermal cost of reheating, 280-284 
Thermodynamic laws, $6 et seq. 
Thomson-Houston solenoid pick machine, 389 
Three-stage compression, work of, 163, 164 
Three-stage compressors, 96, 474, 476, 480, 483, 485, 488 
Track resistance of mine cars, 472, 473 

Transmission losses, comparison of air and steam, 4; in pipes, 248-257 
Transmission of power by compressed air, 248-260 
Transvaal Inst. Mech. Engs., 378 
Triumph drill, 300, 319-321 

Two-pipe system of compressed-air transmission, 272, 273 
Two-stage compression, work of, 70 et seq., 163, 164 
Tunneling, i, 2, 5, 336, 375 
Turbine wheel for driving compressors, 36 

U 

Undercutting machines for coal, 380 et seq. 

Underground air receivers, 193, 277, 432 

Underground reheaters, 290 

Unloaders for air cyhnders, 199-208; Franklin, 206; Ingersoll-Sergeant, 206-207; 

Laidlaw-Dunn-Gordon, 208; Norwalk, 199; Rand, 202, 203; Sullivan, 204, 

205 
Unwin, W. C, 256, 284 



Valve, adjustable cut-off, for steam, 34 
Valveless drills, 319-321, 344, 348-354, 358-361 



INDEX 501 

Valves, air-inlet, 115-135; area of, 117, 118, 127; of Bailey & Co., 158; chat- 
tering of, 119, 120; clack, 115, 131; Corliss, 116, 142-151; of Dover Iron 
Co., 158; Guttermuth, 131; Humboldt rubber-ring, 130; inertia of, 117- 
119; Ingersoll-Rand " hurricane " inlet, 126-128; Johnson, 129; Koster piston 
valve, 158; Laidlaw-Dunn-Gordon, 118, 119; Lens cam-controlled, 156; 
Leyner flat annular, 131-133; mechanically controlled, 115, 116, 117, 142- 
158, 220; Neumann and Esser, 158; Nordberg, 82, 116, 213; Norwalk, 118; 
Pokorny and Wittekind, 158; requisites of, 116; resistance of, 117, 120- 
123,220; Riedler double seat poppet, 152-156; Schuechtermann and Kremer, 
139; Schuetz, 139; skip-valve, 124, 126; springs for, 117-121; sticking of, 
123; Sturgeon, 157 

Valves, air-deHvery, 136-141; air-cataract, 139; Allis-Chalmers, 149; area of, 140, 
141; Corliss, 139, 143-146; Laidlaw-Dunn-Gordon "Cincinnati" valve 
gear, 148; poppet, 147-15 1; Nordberg, 144, 146; Norwalk, 144, 145; oil- 
cataract, 138; Riedler "express," 139, 140; double-seat poppet, 152-156 

Valves of machine drills, 300 et seq.; 332, 334 et seq. 

Van Nostrand's Science Series No. to6, Unwin, 256 

Vauclain compound compressed-air locomotives, 484 

Vekol Gold Mine, Arizona, drilling record, 339 

Velocity of flow of air in pipes, 232,' 256, 257 

Victoria Copper Mine, Michigan, hydraulic air compressor at, 242-245 

Village Deep Mine, S. A., 374 

Volume of air for non-expansive working pumps, 428-431 

Volumes and pressures of compressed air, 159, 160, 165; at altitudes above sea-level, 
216-220; influence of reheating on, 280-285, 292 

Vulcan Iron Works, 48 

W 

Wabana Iron Mines, N. S., drilling record, 338 

"Wahnvolf " drill, 300 

Wainwright water heater employed as reheater, 436 

Wandsworth (England) test on air-Hft pump, 449 

Water-driven compressors, 2>^ et seq.; at Goleta Mine, 36; at Morning Mine, 40, 
42-44; at North Star Mine, 40, 42, 43 

"Water" Leyner drill, 344, 345, 348 

Water-wheels for driving compressors: Knight, 36; Pelton, 36, 40-42; peripheral 
velocity of, 36; Risdon, 36, 37 

Waugh hammer drill, 366-369, 373, 374, 375 

Webb, R. L., 174 

Weber, F. C, 270 

Weight and volume of dry air, 50 

Wet compressors, 62, 63, 75-80, 90-94 

"Wet" reheating, 292, 293 

Weymouth, C. R., 293 

Whitcomb hammer drill, 371 

Wilson ColHery, Pa., compressed-air haulage at, 483 

Wilson tests on air-lift pumps, 450-453 

Winstanley coal cutter, 388 

Wolverine Copper Mine, Mich., drilling record, 340 

Wonder hammer drill, 348, 374 

Woodbridge, D. E., 244 

Work done by air engines, 265-270 

Work done by air-Hft pumps, 448-450, 452 

Work gained by reheating, 281-284 

Work required to compress air, 56 et seq.; 92, no, 159 et seq.; in stage compres- 
sion, 163, 164 

Worthington compound pump at Gwin Mine, 436 

Worthington, Henry R., 423 



502 



INDEX 



Yakima, North, Wash., tunnels, drilling record, 336 
Yoch pick machine, 389 
Yorkshire coal cutter, 388 



Zahner, "Transmission of Power by Compressed Air," 53, 78, 70, 264 
Zeitschrift jilr das Berg-, Hutten-, und Salinen-Wesen, 193 



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